Thermal management systems

ABSTRACT

An evaporator includes a cold plate configured to extract heat from one or more heat loads in proximity to the evaporator. The cold plate includes a housing, and a plurality of channels disposed through the housing, with at least one of the plurality of channels being a meandered channel.

CLAIM OF PRIORITY

This application claims priority under 35 U.S.C. § 119(e) to U.S. Provisional Patent Application Ser. No. 63/214,334, filed on Jun. 24, 2021, and entitled “THERMAL MANAGEMENT SYSTEMS,” the entire contents of which are hereby incorporated by reference.

BACKGROUND

Refrigeration systems absorb thermal energy from heat sources operating at temperatures below the temperature of the surrounding environment and discharge thermal energy into the surrounding environment. Heat sources operating at temperatures above the surrounding environment can be naturally cooled by the surrounding if there is direct contact between the source and the environment.

Conventional refrigeration systems include a compressor, a heat rejection exchanger (i.e., a condenser), a receiver, an expansion device, and a heat absorption exchanger (i.e., an evaporator). Such systems can be used to maintain operating temperature set points for a wide variety of cooled heat sources (loads, processes, equipment, systems) thermally interacting with the evaporator. Closed-circuit refrigeration systems may pump significant amounts of absorbed thermal energy from heat sources into the surrounding environment.

In closed-circuit systems, compressors are used to compress vapor from an evaporating pressure the evaporator and to a condensing pressure in the condensers and condense the compressed vapor converting the vapor into a liquid at a temperature higher than the surrounding environment temperature. The combination of condensers and compressors can add a significant amount of weight and can consume relatively large amounts of electrical power. In general, the larger the amount of absorbed thermal energy that the system is designed to handle, the heavier the refrigeration system and the larger the amount of power consumed during operation, even when cooling of a heat source occurs over relatively short time periods.

In some cases the surrounding environment temperature can appear below the heat source temperature. The refrigeration system provides a contact via refrigerant. There may be no need to compress vapor from the evaporating to condensing pressure since condensation can happen at a pressure slightly higher or even below the evaporating pressure.

SUMMARY

In an example implementation, an evaporator includes a cold plate configured to extract heat from one or more heat loads in proximity to the evaporator. The cold plate includes a housing, and a plurality of channels disposed through the housing, with at least one of the plurality of channels being a meandered channel.

In an aspect combinable with the example implementation, the meandered channel has at least one curved region in the meandered channel.

In another aspect combinable with any of the previous aspects, the meandered channel has one curved region in the meandered channel and the meandered channel has a net internal angle greater than zero degrees and up to 360 degrees.

In another aspect combinable with any of the previous aspects, the meandered channel has more than one curved region in the meandered channel and the meandered channel has a net internal angle between and including zero degrees up to but not including 360 degrees.

In another aspect combinable with any of the previous aspects, the at least one curved region in the meandered channel is defined as a first section of the channel that forms a non-zero degree angle along the length of the channel with respect to a second section of the channel.

In another aspect combinable with any of the previous aspects, the cold plate is a first cold plate and the evaporator includes a plurality of cold plates including the first cold plate.

In another aspect combinable with any of the previous aspects, each of the plurality of cold plates includes at least one meandered channel.

In another aspect combinable with any of the previous aspects, at least one of the plurality of cold plates includes the at least one meandered channel.

Another aspect combinable with any of the previous aspects further includes a plurality of inlet headers fluidly coupled to inlets of the plurality of cold plates; and a plurality of exit headers fluidly coupled to outlets of the plurality of cold plates.

Another aspect combinable with any of the previous aspects further includes an inlet distributer that includes an inlet configured to receive refrigerant fluid; and a plurality of outlets, each of the plurality of outlets fluidly coupled to a corresponding one of the plurality of inlet headers.

Another aspect combinable with any of the previous aspects further includes an exit collector having a plurality of inlets configured to receive refrigerant fluid from the plurality of exit headers.

In another example implementation, a thermal management system includes an evaporator that includes at least one cold plate configured to extract heat from one or more heat loads in proximity to the evaporator. The cold plate includes a housing and a plurality of channels disposed through the housing, with at least one of the plurality of channels being a meandered channel. The system further includes a receiver configured to store refrigerant fluid, the receiver disposed in a refrigerant fluid path with the evaporator; and an expansion valve positioned between the receiver and the evaporator in the refrigerant fluid path. The expansion valve is configured to expand the refrigerant fluid from the receiver and deliver a mixed liquid/vapor refrigerant to the evaporator.

In an aspect combinable with the example implementation, the meandered channel includes at least one curved region in the meandered channel.

In another aspect combinable with any of the previous aspects, the meandered channel includes one curved region in the meandered channel and the meandered channel has a net internal angle greater than zero degrees up to 360 degrees.

In another aspect combinable with any of the previous aspects, the at least one curved region in the meandered channel is defined as a first section of the channel that forms a non-zero degree angle along the length of the channel with respect to a second section of the channel.

In another aspect combinable with any of the previous aspects, the cold plate is a first cold plate and the evaporator includes a plurality of cold plates including the first cold plate and each of the plurality of cold plates includes at least one meandered channel.

In another aspect combinable with any of the previous aspects, the evaporator further includes a plurality of inlet headers coupled to inlets of the plurality of cold plates; and a plurality of exit headers, coupled to outlets of the plurality of cold plates.

In another aspect combinable with any of the previous aspects, the evaporator further includes an inlet distributer including an inlet configured to receive the refrigerant fluid and a plurality of outlets, each of the plurality of outlets fluidly coupled to a corresponding one of the plurality of inlet headers; and an exit collector including a plurality of inlets configured to receive the refrigerant fluid from the plurality of exit headers.

In another aspect combinable with any of the previous aspects, the refrigerant fluid path is an open-circuit refrigerant fluid path.

Another aspect combinable with any of the previous aspects further includes a back-pressure regulator coupled to an outlet of the evaporator and configured to control vapor pressure upstream of the back-pressure regulator.

In another aspect combinable with any of the previous aspects, the refrigerant fluid path further includes a closed-circuit refrigerant fluid path.

Another aspect combinable with any of the previous aspects further includes a compressor configured to receive refrigerant vapor from the evaporator and compress the refrigerant vapor; and a condenser configured to condense the compressed refrigerant vapor from the compressor.

In another aspect combinable with any of the previous aspects, the expansion valve is configured to perform a constant-enthalpy expansion of the liquid refrigerant fluid to generate the refrigerant fluid mixture.

In another aspect combinable with any of the previous aspects, the refrigerant fluid includes ammonia.

In another example implementation, a method of operating an evaporator includes transporting refrigerant fluid through a plurality of channels in a cold plate according to a required mass velocity range and a required mass flow rate demand, with at least one channel of the plurality of channels being a meandered channel; transferring refrigerant through the plurality of channels to absorb an amount of heat from a heat load to provide at an evaporator outlet a preset exit vapor quality; and maintaining a heat load temperature band of the heat load, while thermally conducting the heat from the heat load to the plurality of refrigerant channels within the cold plate.

In an aspect combinable with the example implementation, the meandered channel has at least one curved region in the meandered channel.

In another aspect combinable with any of the previous aspects, the meandered channel has one curved region in the meandered channel and the meandered channel has a net internal angle greater than zero degrees up to 360 degrees.

In another aspect combinable with any of the previous aspects, the at least one curved region in the meandered channel is defined as a first section of the channel that forms a non-zero degree angle along the length of the channel with respect to a second section of the channel.

One or more of the above aspects or other of the disclosed aspects may include one or more of the following advantages. The aspects enable cooling of large heat loads and high heat flux that are also highly temperature sensitive with evaporators specifically configured for two-phase cooling, i.e., cooling heat loads with a liquid-vapor phase refrigerant mixture, which overcomes many of the limitations of evaporators used in conventional closed-cycle refrigeration systems.

The above aspects address at least some of the challenges encountered when using cold plates operating with high heat flux with two-phase refrigerant cooling. The aspects enable control of the exit vapor quality to be below the critical vapor quality and preferably below the mist and dryout regions of the refrigerant phases in order to maintain high boiling heat transfer coefficient. The aspects enable even refrigerant distribution of liquid and vapor phases in inlet headers of the evaporator. The aspects provide for adequate heat loading of each channel of an evaporator to enable a maximal vapor quality at each channel exit and to minimize mass flow rate demand. The aspects enable adequate thermal conduction passes within cold plates from the heat loads to the refrigerant channels to provide an adequate temperature band for all heat loads.

The disclosed evaporators effectively extract heat from the heat loads. Ideally, the channels are built in such a way that each channel receives heat directly from the heat loads and indirectly via axial thermal conductivity causing evaporation of the liquid refrigerant portion to achieve the vapor quality equal to the critical vapor quality and preferably below the mist and dryout regions of the refrigerant phases at the evaporator outlet. This enables evaporation of the maximal amount of liquid refrigerant while preferably avoiding the mist and dryout regions of the refrigerant phases. The inlet header facilitates distribution of liquid and vapor refrigerant to enable such evaporation of the liquid refrigerant. If all channels are equivalent and the thermal loading is equivalent, then an even distribution of the liquid and vapor portion is provided. In some embodiments, providing equivalent channels and/or equivalent thermal loading is not possible or does not make sense (for example, when the lengths of the channels cannot be the same), then in those embodiments the inlet header provides adequate distribution to address the above mentioned challenges.

The details of one or more embodiments are set forth in the accompanying drawings and the description below. Other features and advantages will be apparent from the description, drawings, and claims.

DESCRIPTION OF DRAWINGS

FIG. 1 is a schematic diagram of an example of a thermal management system (TMS) that includes an open-circuit refrigeration system (OCRS) and a cold plate according to the present disclosure.

FIG. 2 is a schematic diagram of an example implementation of a cold plate for a TMS according to the present disclosure.

FIGS. 3A-3C are top views of different example implementations of cold plates with different channel arrangements according to the present disclosure.

FIGS. 4A and 4B are cross-sectional views of different example implementations of cold plates with different channel heights, relative to a cold plate height, according to the present disclosure.

FIGS. 5A and 5B are schematic diagrams that depict different refrigerant distribution feeding mechanisms for cold plates according to the present disclosure.

FIG. 6 is a plot of mass velocity v. vapor quality of two-phase flow patterns.

FIG. 7 is a plot of heat transfer coefficient v. vapor quality.

FIGS. 8-10 are schematic diagrams of other examples implementations of a TMS that includes an OCRS fluidly coupled to a closed-circuit refrigeration system (CCRS) and one or more cold plates according to the present disclosure.

FIG. 11 is a block diagram of a control system or controller according to the present disclosure.

FIG. 12 is a schematic diagram of an example implementations of a TMS that includes a power generation apparatus according to the present disclosure.

FIG. 13 is a schematic diagram of an example of directed energy system that includes a TMS according to the present disclosure.

DETAILED DESCRIPTION

Cooling of large loads and high heat flux that are also highly temperature sensitive can present a number of challenges. On one hand, such loads generate significant quantities of heat that is extracted during cooling. In conventional closed-cycle refrigeration systems, cooling high heat flux typically involves circulating refrigerant fluid at a relatively high mass flow rate. However, closed-cycle system components that are used for refrigerant fluid circulation—including large compressors to compress vapor at a low pressure to vapor at a high pressure and condensers to remove heat from the compressed vapor at the high pressure and convert to a liquid—are heavy and consume significant power. As a result, many closed-cycle systems are not well suited for deployment in mobile platforms—such as on small vehicles or in space—where size and weight constraints may make the use of large compressors and condensers impractical.

On the other hand, temperature sensitive loads such as electronic components and devices may require temperature regulation within a relatively narrow range of operating temperatures. Maintaining the temperature of such a load to within a small tolerance of a temperature set point can be challenging when a single-phase refrigerant fluid is used for heat extraction, since the refrigerant fluid itself will increase in temperature as heat is absorbed from the load.

Directed energy systems that are mounted to mobile vehicles, such as trucks, or that exist in space may present many of the foregoing operating challenges, as such systems may include high heat flux and temperature sensitive components that require precise cooling during operation in a relatively short time. The thermal management systems disclosed herein, while generally applicable to the cooling of a wide variety of heat loads, are particularly well suited for operation with such directed energy systems.

In some cases, the TMS may be specified to cool two different kinds of heat loads—high heat loads (high heat flux, highly temperature sensitive components) operative for short periods of time and low heat loads (relative to the high heat loads) operative continuously or for relatively long periods (relative to the high heat loads). However, to specify a refrigeration system for the high heat load may result in a relatively large and heavy refrigeration system with a concomitant need for a large and heavy power system to sustain operation of the refrigeration system.

Such systems may not be acceptable for mobile applications. Also, start-up and/or transient processes may exceed the short period in which cooling duty is applied for the high heat loads that are operative for short periods of time. Transient operation of such systems cannot provide precise temperature control. Therefore, thermal energy storage (TES) units are integrated with small refrigeration systems and recharging of such TES units are used instead. Still TES units may be too heavy and too large for mobile applications. In addition, such systems are complex devices and reliability may present problems especially for critical applications.

For cooling of large heat loads and high heat flux that are also highly temperature sensitive using conventional refrigeration systems would require use of closed-cycle system components such as relatively large and heavy compressors to compress vapor at a low pressure to vapor at a high pressure and relatively large and heavy condensers to remove heat from the compressed vapor. In addition to being large and heavy, these components typically consume significant amounts of electrical power.

One solution to these problems is the use of two-phase cooling. The primary goal of two-phase cooling is to absorb heat from the heat loads and satisfy certain temperature tolerances and bands. Two-phase cooling is the most effective way to cool electrical components having tight temperature tolerances and narrow temperature bands. The evaporator is a crucial part of any two-phase cooling architecture. The evaporator typically includes one or more cold plates, and, in the present disclosure, such terms may be used interchangeably unless otherwise notes.

There are at least four major challenges in configuring cold plates operating with high heat flux. One challenge involves providing control of an exit vapor quality that is below a critical vapor quality in order to maintain a high boiling heat transfer coefficient and effectively evaporate maximal possible amount of liquid refrigerant. Another challenge involves providing control of an exit vapor quality that is below the critical vapor quality and also below dryout and mist regions of refrigerant phases in order to maintain an even higher boiling heat transfer coefficient and to make maximal possible use of the liquid refrigerant.

Another challenge is to provide even or adequate refrigerant distribution of liquid and vapor phases in inlet headers of an evaporator to provide a sufficient amount of refrigerant for each channel of a multi-channel evaporator. Another challenge is to provide for adequate heat loading of each channel of an evaporator to enable maximal, i.e., critical vapor quality but that is below dryout and mist regions at each channel exit and to minimize mass flow rate demand. Another challenge is to provide adequate thermal conduction passes within cold plates from the heat loads to the refrigerant channels to provide an adequate temperature band for all heat loads with minimal numbers of channels and/or minimal total length of the refrigerant channels.

In particular, the thermal management systems and methods disclosed herein include a number of features that reduce both overall size and weight relative to conventional refrigeration systems, and still extract excess heat energy from both high heat flux, highly temperature sensitive components and relatively temperature insensitive components, to accurately match temperature set points for the components.

At the same time, the disclosed thermal management systems are undersized for the given application. That is, the refrigerant flow rate is less than the required refrigerant for a given amount of refrigeration over a specified period(s) of operation. Whereas conventional refrigeration systems would be designed for the maximum flow rate needed for refrigeration, the systems and methods disclosed herein use a modified evaporator design that maintains even refrigerant distribution to refrigerate heat loads and maintain temperature of those heat loads with defined temperature ranges.

In some aspects, “refrigeration” as used in the present disclosure can mean a system (or multiple systems fluidly coupled) that operates to generate a purposeful change of a characteristic of a coolant (e.g., a refrigerant fluid) to effectuate or increase heat transfer between two mediums (one of which can be the coolant). The purposeful change of the characteristic can be, for example, a change in pressure (e.g., depressurization) of a pressurized coolant though an expansion valve. In some embodiments, the change in pressure can include a phase change of the coolant, such as a liquid-to-gas phase change (e.g., endothermic vaporization). In some embodiments, pressurization of the refrigerant can be performed by a powered (e.g., electrically or otherwise) component, such as (but not limited to) a compressor. In some embodiments, pressurization can be performed as part of the refrigeration cycle (e.g., a closed-cycle refrigeration process in which gaseous refrigerant is substantially or completely recycled and compressed into a liquid state) or prior to use (e.g., storing pre-compressed liquid refrigerant for later use in an open-cycle refrigeration process in which a reserve of liquid refrigerant is used but substantially not recycled). In some embodiments, the phase change can be driven by heating a liquid refrigerant with a very low boiling point (e.g., ammonia as used in an absorption-type refrigeration cycle).

Referring now to FIG. 1 , a thermal management system (TMS) 100 includes an open circuit refrigeration system (OCRS) 5 that utilizes one or more cold plates 116 to cool one or more heat loads 118. The OCRS 5 includes a receiver 110 having a receiver inlet 109 and receiver outlet 111, an optional solenoid valve 112 having an inlet 113 and an outlet 115, a flow control device 114, an evaporator 116 especially adapted for two-phase operation, i.e., a two-phase evaporator 116 also called a cold plate 116 having an inlet 121 and an outlet 123, and a solenoid valve 112 having an inlet 17 and an outlet 19. The aforementioned components of OCRS 5 are interconnected by one or more conduit sections to form an open circuit refrigerant fluid path.

In this example, one or more heat loads 118 can be considered high heat loads that are in thermal conductive and/or convective contact or in proximity with the cold plate 116. OCRS 5 may optionally include a gas receiver 10 with an outlet 11 fluidly coupled to the inlet 109 of the receiver 110 via conduit, such that a gas flow path extends between the gas receiver 10 and the receiver 110 (that stores refrigerant fluid 1). An optional flow control device 12 having inlet 13 and outlet 15, as well as an optional check valve 14 are positioned along the gas flow path between the optional gas receiver 10 and the receiver 110.

Receiver 110 is typically implemented as an insulated vessel that stores a refrigerant fluid at relatively high pressure. When ambient temperature is very low and, as a result, pressure in the receiver 110 is low and insufficient to drive refrigerant fluid flow through the TMS 100, gas from gas receiver 10 can be directed into receiver 110. The gas compresses liquid refrigerant fluid 1 in receiver 110, maintaining the liquid refrigerant fluid 1 in a sub-cooled state, even when the ambient temperature and the temperature of the liquid refrigerant fluid are relatively high. Receiver 110 can also include insulation applied around the receiver 110 and a heater to reduce thermal losses.

In some aspects, receiver 110 includes the inlet 109 and the outlet 111, and may include an optional pressure relief valve. To charge receiver 110, refrigerant fluid is typically introduced into receiver 110 via the inlet 109, and this can be done, for example, at service locations. Operating in the field the refrigerant exits receiver 110 through outlet 111 that is connected to conduit. In case of emergency, if the fluid pressure within receiver 110 exceeds a pressure limit value, a pressure relief valve opens to allow a portion of the refrigerant fluid to escape through valve to reduce the fluid pressure within receiver 110. Receiver 110 is typically implemented as an insulated vessel that stores a refrigerant fluid at relatively high pressure. Receiver 110 can also include insulation applied around the receiver to reduce thermal losses.

In general, receiver 110 can have a variety of different shapes. In some embodiments, for example, the receiver is cylindrical. Examples of other possible shapes include, but are not limited to, rectangular prismatic, cubic, and conical. In certain embodiments, receiver 110 can be oriented such that outlet 111 is positioned at the bottom of the receiver 110. In this manner, the liquid portion of the refrigerant fluid 1 within receiver 110 is discharged first through outlet 111, prior to discharge of refrigerant vapor. In certain embodiments, the refrigerant fluid 1 can be an ammonia-based mixture that includes ammonia and one or more other substances. For example, mixtures can include one or more additives that facilitate ammonia absorption or ammonia burning.

The cold plate 116 can be implemented in a variety of ways. In general, cold plate 116 functions as a heat exchanger, providing thermal contact between the refrigerant fluid and the high heat load 118. Typically, cold plate 116 allows streams of refrigerant fluid to flow through the cold plate 116 and absorb heat from the high heat load 118. In some aspects, light weight and compactness are generally requirements for TMS systems to cool electronic components that are mobile or space based. In some cases, cold plates and/or heat loads are inherently heavy and can carry substantial amount of thermal inertia. The cold plate 116 discussed herein uses this built-in thermal inertia to increase the TMS cooling effectiveness and maintain the temperature tolerances in the range when undersized cooling systems are used.

Turning now to FIG. 2 , this figure shows an example implementation of a cold plate 200 that can be used as the cold plate 116 in the present disclosure. In this example, cold plate 200 can be a mini-channel tube that is directly attached to a heat load such as high heat load 118 or other heat loads according to the present disclosure. Shown in FIG. 2 is the cold plate 200 with a heat exchanging core 204 and multiple parallel channels 202 of equal length, a short heat conduction pass between the heat load 118 and channels 202, and a large cross section of the heat conduction pass. The cross-section shape of the channels 202 is round, but the shape can be configured to be square, rectangular, triangular, etc. FIG. 2 shows the heat load 118 attached to one side of the cold plate 200, but alternative embodiments may have heat loads on both sides of the cold plate 200. The cold plate 200 includes an inlet header 206 and an exit header 208 (and thus, multiple cold plates 200 could include multiple inlet headers 206 and multiple exit headers 208). The refrigerant channels 202 are routed inside the cold plate 200 between the headers 206 and 208.

In a modified implementation of cold plate 200, more robust channel routing arrangements can be used to mitigate the challenge of uneven refrigerant distribution of liquid and vapor phases between the evaporator channels. The channel routing arrangements should depend on locations of the heat loads on a heat map, including the heat loads, heat flux, heat load operating temperatures, and heat load temperature bands. The total heat load dictates the mass flow rate. The channel routing arrangements discussed below take into consideration maintaining adequate mass velocity (flux) range addressing the challenge of FIG. 2 . The mass flow rate demand dictates the number of channels and their cross-section dimensions. The length of each channel 202 is dictated by the heat load, required heat transfer area, and refrigerant pressure drop. The channels 202 cross a combination of heat loads in order to absorb a specified amount of heat to reach an exit vapor quality and maintain proper heat load temperature band of all heat loads. The cold plate 200 may allow superheated region if the superheated region would maintain the heat load to be within the allowed temperature band. The channel routing arrangement also takes into consideration maintaining adequate thermal conduction from the heat loads to the refrigerant channels within cold plates to provide an adequate temperature band for heat loads have not been crossed by the refrigerant channels and the refrigerant routing of the refrigerant channels should not interfere with heat load mounting arrangements.

The cold plate 200 may have various heat load distributions (heat maps), which require a more complicated geometry of cooling refrigerant channels/circuits. To properly handle the various heat load distributions (heat maps) the channels 202 may have different lengths, different cross-section geometry, different shapes, but all of them are configured to generate an adequate thermodynamic exit states to enable adequate loading of each channel/circuit and operation of the heat loads, within the allowed temperature band.

Turning now to FIGS. 3A-3C, these figures show example implementations of other cold plates that can be used as the cold plate 116 in the present disclosure. Each of the implementations has different locations for inlet and exit headers. Each of the cold plates in FIGS. 3A-3C are configured to have an inlet header and an exit header coupled to a plurality of channels. A refrigerant distributor (such as refrigerant distributor 504 discussed later) can be used to distribute refrigerant into the inlet header(s), while a refrigerant collector can be used to collect refrigerant from the outlet header(s).

FIG. 3A shows an inlet header 306 and an exit header 308 at a pair of opposing walls of a housing 302 for a cold plate 300. In FIG. 3A, channels 304 are disposed between the inlet header 306 and exit header 308. Two channels 304 as shown are meandered channels that are examples of channels that wander or roam along the cold plate 300 between the inlet header 306 and exit header 308. One non-meandered channel 304 is shown that connects in a straight line between the inlet header 306 and exit header 306. Meandered channels and non-meandered channel have channel paths that correspond to a heat distribution (heat map) of loads carried by the cold plate 300.

FIG. 3B shows the inlet header 326 and the exit header 328 disposed at a pair of adjacent walls of a housing 322 of a cold plate 320. Two example meandered channels 324 are disposed between the inlet header 326 and exit header 328 in a meandering arrangement, while one example non-meandered channel 324 connects in a straight line between the inlet header 326 and exit header 328. The meandered channels have one or more bends or curves between the inlet header 326 and the exit header 328 that have a sum of internal angular bends between the inlet header 326 and the exit header 328 that are approximately 90 degrees. Meandered channels and non-meandered channel have channel paths that correspond to a heat distribution (heat map) of loads carried by the cold plate 320.

FIG. 3C shows inlet header 346 and exit header 328 disposed on the same one of the walls of a housing 342 of a cold plate 340. Meandered channels 344 are disposed between the inlet header 346 and exit header 348 in a generally meandering arrangement. The meandered channels 344 have one or more bends or curves between the inlet header 346 and the exit header 348 that have a sum of internal angular bends between the inlet header 346 and the exit header 348 that are approximately 180 degrees. In some aspects, a meandered channel can be a crossover channel that is disposed between the inlet header 346 and exit header 348 and crosses over a portion of the meandered channels. Meandered channels and crossover meandered channels have channel paths that correspond to a heat distribution (heat map) of loads carried by the cold plate 340.

FIGS. 3A-3C each depict a single cold plate 300 or 320 or 340. However, in some embodiments, the single cold plates 300 or 320 or 340 can be a plurality of the single cold plates 300 or 320 or 340. In some embodiments, the single cold plate 300 or 320 or 340 can be intermixed, i.e., an embodiment could have one or more of the cold plates 300 and one or more of the cold plates 320. Another embodiment could have one or more of the cold plates 300 and one or more of the cold plates 340. All refrigerant headers may be engaged simultaneously. Alternatively, combinations of inlet headers, related refrigerant channels, and exit headers may be engaged for operating at specific operating conditions.

In some aspects, a meandered channel is a channel in the cold plate that has a non-straight (e.g., curved) fluid path between inlet and outlet of the channel in the cold plate. A meandered channel may have a net internal angle between 0 degrees and 360 degrees. By net is meant the modulo angle at which the inlet of the channel lies relative to the outlet of the channel. A meandered channel may have at least one curve in the meandered channel that provides a net internal angle greater than 0 degrees up to 360 degrees. A meandered channel may have two curves (or more) that provide a net 0 degree angle between the inlet of the channel relative to the outlet of the channel or any net internal angle between 0 degrees and 360 degrees. A curve in a meandered channel is defined as a first section of the channel that forms a non-zero degree angle along the length of the channel with respect to a second section of the channel, i.e., a bend between the first section and the second section of the channel.

A channel routing arrangement of the channels in the cold plate(s) is governed by the heat map corresponding to the load. In ideal cases the heat loads are evenly placed on the cold plates. In actual implementations, often the heat loads are not evenly placed on the cold plates. For example, it could occur that some areas of cold plates are not loaded at all and there are some areas where the most heat loads are concentrated. In order to allocate a large plurality of mini-channels in the region with the concentrated heat loads, the multiple layers of mini/micro-channels should be arranged.

Turning now to FIGS. 4A-4C, these figures show example implementations of other cold plates 400 and 420 that can be used as the cold plate 116 in the present disclosure. As shown, different allocations of channels (generally) 404 and 424 are shown. FIG. 4A shows a conventional allocation of the channels 404 in the cross-section of a housing 402 of the cold plate 400. FIG. 4B shows an allocation of the channels 424 in the cross-section of a housing 402 of the cold plate 400 with respect to a height of cold plate 420. By placing channels 424 at different heights within the cold plate 420, and further by allowing the channels to be in proximity to one another at different heights (one channel 424 being over one or more other channels 424, as shown), can allow for an increase in density of the refrigerant channels and avoids intersecting the channels in those cold plates having complicated heat maps. The cross-section shape of the channels is round, but the shape can be configured to be square, rectangular, triangular, etc.

Turning now to FIGS. 5A and 5B, these figures show example implementations of other cold plate systems that can integrate multiple cold plates that can be used as the cold plate 116 in TMS 100 (and other TMS according to the present disclosure). FIG. 5A shows a cold plate assembly 500 that integrates multiple cold plates 502. The assembly 500 includes a refrigerant distributor 504 having conduit(s) 506 that feed each inlet header 508 that is attached to cold plates 502. FIG. 8B shows a cold plate assembly 550 that integrates multiple cold plates 552. The assembly 550 includes a refrigerant distributor 554 with conduits 556 feeding a plurality of inlet headers 558 configured to feed a plural refrigerant channels (not shown) disposed in the cold plates 552. Alternatively, as shown, each inlet header may be fed by an individual expansion device 114. As with inlet headers, there can be a plurality of outlet headers (not shown) in the assembly 550. If an ejector is used, a common or individual liquid separator may be located at the cold plate outlet (ejectors and liquid separators discussed later).

A variety of inlet header configurations can be used. For example, longitudinal configurations that are oriented vertically and horizontally, as well as radial headers can be used. Examples are discussed in Design Tradeoffs in Microchannel Heat Exchangers, by T. Kulkarni and C. W. Bullard, ACRC TR-208 Feb. 1, 2003). The inlet (horizontal and vertical) headers of the cold plates are configured to maintain even refrigerant distribution of liquid and vapor phases through the channels. The inlet headers may be configured as stand-alone components, or they may be integrated with the cold plate. The inlet (horizontal and vertical) headers of the cold plates are configured to maintain even refrigerant distribution of both liquid and vapor phases. Header features such as header geometry, header orientation, tube protrusion, fluid properties, inlet mass flux and vapor quality are controlled in order to provide even refrigerant distribution.

Cold plate 116, and the example cold plates described previously, are used as two-phase evaporators as noted. Two-phase flow regime is characterized by different flow patterns. In a paper by Wojtan L, Ursenbacher T, Thome J. R. 12005 (Investigation of flow boiling in horizontal tubes, Part I Int. J. Heat Mass Transfer, 48, 2955-2969) the authors review a wide variety of flow channels and conditions. FIG. 6 shows a plot 600 of a two-phase flow pattern from that paper. The plot 80 shows intermittent (I), annual (A), slug (S), stratified (S), stratified wavy (SW), dry-out (D), and mist (M) flow regimes in mass velocity v. vapor quality coordinates. A border line 81 on the left of the dry-out flow regime represents so-called critical vapor quality states. The two-phase flow pattern maps are different for different refrigerants. All boundary curve shapes shown on the map change with changing operating conditions, such as pressure, temperature, mass velocity, adiabatic or diabatic flow, and channel orientation.

FIG. 7 is another plot 700 from a different paper by Wojtan L, Ursenbacher T, Thome J. R. 12005 (Investigation of flow boiling in horizontal tubes, Part II Int. J. Heat Mass Transfer, 48, 2970-2985). This plot 700 shows a boiling heat transfer coefficient behavior for the same operating conditions as in the plot on FIG. 6 . FIG. 7 shows very low boiling heat transfer coefficients in the dry-out and mixed flow regimes 702 compared with those in the other flow regimes. These very low boiling heat transfer coefficients indicate that the dry-out and mist flow regimes should be avoided in evaporators dealing with high heat flux heat loads. Therefore, the first challenge for cold plates operating with high heat flux is that the cold plates are configured to operate with exit vapor qualities below the critical vapor quality.

As used herein a channel is a passage through a cold plate, which is configured to carry refrigerant fluid so as to remove heat from the cold plate. A mini-channel heat exchanger is so called because the channels are of relatively small diameter, and are distributed through the cold plate, in a generally uniform manner. These so-called mini-channel heat exchanging geometries significantly improve boiling heat transfer coefficients. The principles discussed here are also applicable to so called micro-channels where the micro-channels are of even smaller diameters, and macro-channel where the macro-channels are of larger diameters than the micro-channels and the mini-channels.

The mini-channel evaporators, however, produce a second challenge—uneven refrigerant distribution of liquid and vapor phases between the evaporator channels. Generally, the vapor quality at the inlet may vary from zero to 0.2. The higher the vapor quality, the larger the difference between the liquid and vapor density, and the more challenging becomes refrigerant distribution. Therefore, the second challenge, the inlet headers of the cold plates operating with high heat flux heat loads is to enable even refrigerant distribution of liquid and vapor phases between the evaporator channels. Usually it means equal vapor quality and mass flow rate at the evaporator inlet (or inlet header exit).

The third challenge involves adequate heat loading of each channel to enable maximal vapor quality at each channel exit provided that the exit vapor quality is below the critical vapor quality. Usually this means equal loading of refrigerant channels of the same length and shape. However, some heat load arrangements may require configuring cold plates having different lengths, channel shapes, non-equal heat loading and configuring inlet headers to provide non-equal mass flow rate distribution. Adequate heat loading enabling a maximal exit vapor quality that is below the critical vapor quality is still the primary objective for all configurations.

The fourth challenge is to configure adequate thermal conduction passes within the cold plates from the heat loads to the refrigerant channels to provide an adequate temperature band for all heat loads. Thermal resistances on thermal conduction passes between the heat loads and the refrigerant channels will increase the temperature band. Each conduction pass should have an appropriate combination of cross-section area and the length. The cold plate is comprised of a high thermally conductivity material (such as aluminum). All contacted resistances are minimized and/or are equal. Usually the thicker plate is the better in the thermal conductivity, but this makes the cold plate heavier.

The maximal possible vapor quality allows to reduce the refrigerant mass flow rate demand for any given combination of heat loads. The mass flow rate is given by:

$\overset{˙}{m} = \frac{Q}{\Delta h}$

where

—is the mass flow rate, Q is the heat load, Δh=h_(o)−h_(i), with h_(i) being the inlet enthalpy at the evaporator inlet as function of the inlet pressure and vapor quality and h_(o) being the exit enthalpy at the evaporator inlet as function of the exit pressure and vapor quality. A low exit vapor quality wastes portion of liquid refrigerant due to incomplete evaporation. On the other hand, excessive vapor quality leads to abrupt heat transfer rate reduction rate in the dry-out region when the vapor quality is above the critical vapor quality.

The cold plate 116, as described herein, may be part of a TMS that implements any OCRS or CCRS. In some aspects, a TMS may include an individual liquid separator at the evaporator inlet, individual liquid separator at the evaporator outlet, and/or expansion devices. Exit expansion devices include electronic expansion valves, mechanical expansion valves, fixed orifice expansion devices, capillary tubes, or an ejector, as well as a back-pressure regulators, as discussed below. In some cases, specific terms such as “expansion valve” will be used to refer to electronic expansion valves, mechanical expansion valves, and fixed orifice expansion devices, while “ejector” or “back-pressure regulator” are used to refer to those specific components.

Turning back to FIG. 1 , during operation of OCRS 5, cooling can be initiated by a variety of different mechanisms. In some embodiments, for example, OCRS 5 includes a temperature sensor attached to the high heat load 118 (as will be discussed subsequently). When the temperature of the high heat load 118 exceeds a certain temperature set point (i.e., threshold value), a control system 999 (or controller 999) connected to the temperature sensor can initiate cooling of the high heat load 118. Alternatively, in certain embodiments, OCRS 5 operates essentially continuously—provided that the pressure within receiver 110 is sufficient—to cool high heat load 118. As soon as receiver 110 is charged with refrigerant fluid 1, refrigerant fluid 1 is ready to be directed into cold plate 116 to cool high heat load 118.

The TMS 100, as all disclosed embodiments, may also include a control system (or controller) 999 (see FIG. 11 for an exemplary embodiment) that produces control signals (based on sensed thermodynamic properties) to control operation of one or more of the various devices, e.g., optional solenoid control valve 112, expansion valve 114, etc., as needed, as well as to control operation of a motor of the compressor 104, a fan 108, or other components in other example implementations of a TMS. Control system 999 may receive signals, process received signals and send signals (as appropriate) from/to the sensors and control devices to operate the TMS 100.

The term “control system,” as used herein, can refer to an overall system that provides control signals and receives feedback data from unit controllers, such as unit controllers (e.g., programmable logic controllers, motor controllers, variable frequency drives, actuators). In some aspects, the control system includes the overall system and the unit controllers. In some aspects, a control system simply refers to as a single unit controller or a network of two or more individual unit controllers that communicate directly with each other (rather than with an overall system.

The process streams (e.g., refrigerant flows, ambient airflows, other heat exchange fluid flows) in a TMS according to the present disclosure, as well as process streams within any downstream processes with which the TMS is fluidly coupled, can be flowed using one or more flow control systems (e.g., that include the control system 999) implemented throughout the system. A flow control system can include one or more flow pumps, fans, blowers, or solids conveyors to move the process streams, one or more flow pipes through which the process streams are flowed and one or more valves to regulate the flow of streams through the pipes, whether shown in the exemplary figures or not. Each of the configurations described herein can include at least one variable frequency drive (VFD) coupled to a respective pump or fan that is capable of controlling at least one fluid flow rate. In some implementations, liquid flow rates are controlled by at least one flow control valve.

In some embodiments, a flow control system can be operated manually. For example, an operator can set a flow rate for each pump or transfer device and set valve open or close positions to regulate the flow of the process streams through the pipes in the flow control system. Once the operator has set the flow rates and the valve open or close positions for all flow control systems distributed across the system, the flow control system can flow the streams under constant flow conditions, for example, constant volumetric rate or other flow conditions. To change the flow conditions, the operator can manually operate the flow control system, for example, by changing the pump flow rate or the valve open or close position.

In some embodiments, a flow control system can be operated automatically. For example, the flow control system can be connected to a computer or control system (e.g., control system 999) to operate the flow control system. The control system can include a computer-readable medium storing instructions (such as flow control instructions and other instructions) executable by one or more processors to perform operations (such as flow control operations). An operator can set the flow rates and the valve open or close positions for all flow control systems distributed across the facility using the control system. In such embodiments, the operator can manually change the flow conditions by providing inputs through the control system. Also, in such embodiments, the control system can automatically (that is, without manual intervention) control one or more of the flow control systems, for example, using feedback systems connected to the control system. For example, a sensor (such as a pressure sensor, temperature sensor or other sensor) can be connected to a pipe through which a fluid flows. The sensor can monitor and provide a flow condition (such as a pressure, temperature, or other flow condition) of the process stream to the control system. In response to the flow condition exceeding a threshold (such as a threshold pressure value, a threshold temperature value, or other threshold value), the control system can automatically perform operations. For example, if the pressure or temperature in the pipe exceeds the threshold pressure value or the threshold temperature value, respectively, the control system can provide a signal to the pump to decrease a flow rate, a signal to open a valve to relieve the pressure, a signal to shut down process stream flow, or other signals.

In general, cooling is initiated when a user of the TMS 100 or the heat load issues a cooling demand. Upon initiation of a cooling operation, refrigerant fluid 1 from receiver 110 is discharged from outlet 111 and through optional solenoid valve 112 if present. As discussed above, the driving force for the transport of refrigerant fluid through OCRS 5 is the pressure within receiver 110. Refrigerant fluid is transported through conduit to expansion valve 114, which directly or indirectly controls vapor quality (see discussion below) at the evaporator outlet 123. In the following discussion, expansion valve 114 can be any flow control device that performs the functional steps described below and provides for vapor quality control at the evaporator outlet 123.

Once inside the expansion valve 114, the refrigerant fluid undergoes constant enthalpy expansion from an initial pressure p_(r) (i.e., the receiver pressure) to an evaporation pressure p_(e) at the outlet 119 of the expansion valve 114. In general, the evaporation pressure pc depends on a variety of factors, most notably the desired temperature set point value (i.e., the target temperature) at which the high heat load 118 is to be maintained and the heat input generated by the high heat load 118.

The initial pressure in the receiver 110 tends to be in equilibrium with the surrounding temperature and is different for different refrigerant fluids. The pressure in the cold plate 116 depends on the evaporating temperature, which is lower than the heat load temperature and is defined during design of the system. The TMS 100 is operational as long as the receiver-to-evaporator pressure difference is sufficient to drive adequate refrigerant fluid flow through the expansion valve.

After undergoing constant enthalpy expansion in the expansion valve 114, the liquid refrigerant fluid is converted to a mixture of liquid and vapor phases at the temperature of the fluid and evaporation pressure pc. This two-phase refrigerant fluid mixture is transported to inlet 121 of cold plate 116.

When the two-phase mixture of refrigerant fluid is directed into cold plate 116, the liquid phase absorbs heat from high heat load 118, driving a phase transition of the liquid refrigerant fluid into the vapor phase. Because this phase transition occurs at (nominally) constant temperature, the temperature of the refrigerant fluid mixture within cold plate 116 remains nominally unchanged, provided at least some liquid refrigerant fluid remains in cold plate 116 to absorb heat.

Further, the constant temperature of the refrigerant fluid mixture within cold plate 116 can be controlled by adjusting the pressure pc of the refrigerant fluid, since adjustment of p_(e) changes the boiling temperature of the refrigerant fluid. Thus, by regulating the refrigerant fluid pressure pc upstream from cold plate 116 (e.g., using solenoid valve 112), the temperature of the two-phase refrigerant fluid mixture within cold plate 116 (and, nominally, the temperature of high heat load 118) can be controlled to match a specific temperature set-point value for each of the high heat load 118, ensuring that each of the high heat load 118 is maintained at, or very near, its targeted temperature.

The pressure drop across the cold plate 116 causes drop of the temperature of the refrigerant mixture (which is the evaporating temperature), but still the cold plate 116 can be configured to maintain the high heat loads temperatures within set tolerances.

In some embodiments, for example, the evaporation pressure of the refrigerant fluid can be adjusted by solenoid valve 112 to ensure that the temperature of high heat load 118 is maintained to within ±5 degrees C. (e.g., to within ±4 degrees C., to within ±3 degrees C., to within ±2 degrees C., to within ±1 degree C. to within ±0.1 degree C.) of the temperature set point value for high heat load 118.

As discussed above, within cold plate 116, a portion of the liquid refrigerant in the two-phase refrigerant fluid mixture is converted to refrigerant vapor by undergoing a phase change. As a result, the refrigerant fluid mixture that emerges from cold plate 116 has a higher vapor quality (i.e., the fraction of the vapor phase that exists in refrigerant fluid mixture) than the refrigerant fluid mixture that enters cold plate 116.

As the refrigerant fluid mixture emerges from cold plate 116, a portion of the refrigerant fluid can optionally be used to cool one or more additional heat loads (not shown). Typically, for example, the refrigerant fluid that emerges from cold plate 116 is nearly in the vapor phase. The refrigerant fluid vapor (or, more precisely, high vapor quality fluid vapor) can be directed into a heat exchanger (not shown) coupled to another heat load, and can absorb heat from the additional high heat loads during propagation through the heat exchanger.

The two-phase refrigerant fluid mixture emerging from cold plate 116 is transported through conduit to solenoid valve 112, which directly or indirectly controls the upstream pressure, that is, the evaporating pressure pc in the system. After passing through solenoid valve 112, the vapor component of the two-phase refrigerant fluid mixture is discharged as exhaust through an exhaust line 20 of OCRS 5. Refrigerant vapor discharge can occur directly into the environment surrounding OCRS 5 (i.e., in a flow circuit that is “open” to the environment). Alternatively, in some embodiments, the refrigerant vapor can be further processed; various features and aspects of such processing are discussed below.

It should be noted that the foregoing steps, while discussed sequentially for purposes of clarity, occur simultaneously and continuously during cooling operations. In other words, refrigerant vapor is continuously being discharged from receiver 110, refrigerant fluid mixture undergoes continuous expansion in expansion valve 114, flowing continuously through cold plate 116 and solenoid valve 112, and being discharged from OCRS 5, while high heat load 118 is being cooled.

As discussed above, during operation of OCRS 5, as refrigerant fluid is drawn from receiver 110 and used to cool high heat load 118, the pressure driving the refrigerant fluid in receiver 110 through the TMS 100 can be maintained at a constant value for an extended period of operation by introducing gas from gas receiver 10 into receiver 110. In systems where a common receiver is charged with both refrigerant fluid and gas (as described above) or when gas receiver 10 is undercharged initially with gas, the period during which constant pressure can be maintained in receiver 110 may be compromised.

If the pressure within receiver 110 falls sufficiently, the capacity of OCRS 5 to maintain a particular temperature set point value for high heat load 118 may be compromised. Therefore, the pressure in the receiver 110, in the optional gas receiver 10, or the pressure drop across the expansion valve 114 (or any related refrigerant fluid pressure or pressure drop in OCRS 5) can be measured and used to adjust operation of the first expansion device 18.

In addition, one or more measured pressure values can provide an indicator of the remaining operational time. An appropriate warning signal can be issued (e.g., by control system 999) to indicate that, in a certain period of time, the TMS 100 may no longer be able to maintain adequate cooling performance; operation of the TMS 100 can even be halted if the pressure in receiver 110 (or any other measured pressure value in OCRS 5) reaches a low-end threshold value.

It should be noted that while in FIG. 1 only a single receiver 110 is shown, in some embodiments, OCRS 5 can include multiple receivers 110 to allow for operation of the TMS 100 over an extended time period. Each of the multiple receivers can supply refrigerant fluid 1 to the TMS 100 to extend to total operating time period. Some embodiments may include a plurality of cold plates 116 connected in parallel, which may or may not be accompanied by a plurality of expansion valves as well.

The expansion valve 114 functions as a flow control device. In general, expansion valve 114 can be implemented as any one or more of a variety of different mechanical and/or electronic devices. For example, in some embodiments, expansion valve 114 can be implemented as a fixed orifice, a capillary tube, and/or a mechanical or electronic expansion valve. In general, fixed orifices and capillary tubes are passive flow restriction elements which do not actively regulate refrigerant fluid flow.

Mechanical expansion valves (usually called thermostatic or thermal expansion valves) are typically flow control devices that enthalpically expand a refrigerant fluid from a first pressure to an evaporating pressure, controlling the superheat at the evaporator outlet. Mechanical expansion valves generally include an orifice, a moving seat that changes the cross-sectional area of the orifice and the refrigerant fluid volume and mass flow rates, a diaphragm moving the seat, and a bulb at the evaporator outlet. The bulb is charged with a fluid and it hermetically, fluidly communicates with a chamber above the diaphragm. The bulb senses the refrigerant fluid temperature at the evaporator outlet (or another location) and the pressure of the fluid inside the bulb transfers the pressure in the bulb through the chamber to the diaphragm, and moves the diaphragm and the seat to close or to open the orifice.

Typical electrical expansion valves include an orifice, a moving seat, a motor or actuator that changes the position of the seat with respect to the orifice, a controller, and pressure and temperature sensors at the evaporator outlet. The controller calculates the superheat for the expanded refrigerant fluid based on pressure and temperature measurements at the evaporator outlet. If the superheat is above a set-point value, the seat moves to increase the cross-sectional area and the refrigerant fluid volume and mass flow rates to match the superheat set-point value. If the superheat is below the set-point value, the seat moves to decrease the cross-sectional area and the refrigerant fluid flow rates.

Examples of suitable commercially available expansion valves that can function as first expansion valve 114 include, but are not limited to, thermostatic expansion valves available from the Sporlan Division of Parker Hannifin Corporation (Washington, Mo.) and from Danfoss (Syddanmark, Denmark).

The solenoid valve 112 generally functions to control the fluid pressure upstream of the solenoid valve 112. In OCRS 5, solenoid valve 112 controls the refrigerant fluid pressure upstream from the cold plate 116 and solenoid valve 112. In general, solenoid valve 112 can be implemented using a variety of different mechanical and electronic devices. Typically, for example, solenoid valve 112 can be implemented as a flow regulation device, such as a back-pressure regulator. A back-pressure regulator is a device that regulates fluid pressure upstream from the regulator.

In general, a wide range of different mechanical and electrical/electronic devices can be used as solenoid valve 112. Typically, mechanical back-pressure regulating devices have an orifice and a spring supporting the moving seat against the pressure of the refrigerant fluid stream. The moving seat adjusts the cross-sectional area of the orifice and the refrigerant fluid volume and mass flow rates.

Typical electrical back-pressure regulating devices include an orifice, a moving seat, a motor or actuator that changes the position of the seat in respect to the orifice, a controller, and a pressure sensor at the evaporator outlet or at the valve inlet. If the refrigerant fluid pressure is above a set-point value, the seat moves to increase the cross-sectional area of the orifice and the refrigerant fluid volume and mass flow rates to re-establish the set-point pressure value. If the refrigerant fluid pressure is below the set-point value, the seat moves to decrease the cross-sectional area and the refrigerant fluid flow rates.

In general, back-pressure regulators are selected based on the refrigerant fluid volume flow rate, the pressure differential across the regulator, and the pressure and temperature at the regulator inlet. Examples of suitable commercially available back-pressure regulators that can function as solenoid valve 112 include, but are not limited to, valves available from the Sporlan Division of Parker Hannifin Corporation (Washington, Mo.) and from Danfoss (Syddanmark, Denmark).

Flow control device 12 is optional and is positioned between gas receiver 10 and receiver 110. Without the flow control device 12, during operation of OCRS 5, gas in gas receiver 10 is discharged from gas receiver 10 directly into receiver 110 via conduit. When present in OCRS 5, flow control device 12 functions to regulate the pressure within receiver 110, downstream from the flow control device 12. During operation of OCRS 5, flow control device 12 effectively maintains the total pressure within receiver 110 at or above a target pressure value adequate to provide for sub-cooling of refrigerant fluid 1 in receiver 110, which maintains a particular refrigerant mass flow rate through expansion valve 114 and cold plate 116 and, as a result, achieves a desired cooling capacity for one or more high heat loads 118 connected to OCRS 5. If the pressure within receiver 110 falls below the target pressure value, flow control device 12 opens to allow additional gas from gas receiver 10 to enter receiver 110, thereby increasing the pressure within receiver 110.

Flow control device 12 effectively functions as a flow regulation device for the gas in gas receiver 10, and is implemented as any one or more of a variety of different mechanical and/or electronic devices. One example of such a device is a downstream pressure regulator (DPR), which is a device that regulates fluid pressure downstream from the regulator. Examples of suitable commercially available downstream pressure regulators that can function as flow control device 12 include, but are not limited to, regulators available from Emerson Electric (St. Louis, Mo.).

A variety of different gases can be introduced into gas receiver 10 to control the gas pressure in receiver 110. In general, gases that are used are inert (or relatively inert) with respect to the refrigerant fluid. As an example, when a refrigerant fluid 1 such as ammonia is used, suitable gases that can be introduced into gas receiver 10 include, but are not limited to, one or more of nitrogen, argon, xenon, and helium.

Referring now to FIG. 8 , an example implementation of a TMS 800 that includes an OCRS 850 fluidly coupled to a closed-circuit refrigeration system (CCRS) 840 and one or more cold plates 116 is shown. In this example, TMS 800 provides closed-circuit refrigeration for one or more low heat loads 120 over long time intervals and open-circuit refrigeration for refrigeration of one or more high heat loads 118 over short time intervals (relative to the interval of refrigeration of low heat load). More specifically, the TMS 800 includes open-circuit refrigeration system 850, without the optional gas receiver 10 and without the control devices 12 and 16, and further includes a closed-circuit refrigeration system (CCRS) 840.

OCRS 850 (implemented within TMS) includes an open circuit fluid circuit that includes the components/devices of FIG. 1 , that is the receiver 110, the optional solenoid valve 112, the expansion valve 114, the cold plate 116, flow control device 142, exhaust 140, and associated conduit. The high heat load 118 is in thermal conductive and/or convective contact with the cold plate 116. The OCRS 840 also includes a suction accumulator 124 having an inlet 125 and a vapor-side outlet 127. In some embodiments of TMS discussed herein, the suction accumulator 124 is replaced by a liquid separator 124 that also includes a liquid-side outlet (described later). The flow control device 142 is also referred to as back-pressure regulator 142.

The CCRS 840 (implemented within the TMS 800) includes the receiver 110 that includes inlet 109 and outlet 111, the optional solenoid valve 112, the expansion valve 114, the cold plate 116, the suction accumulator 124, a junction 138, a compressor 104 having a compressor inlet 101 and a compressor outlet 103, and a condenser 106 having a condenser inlet 105 and a condenser outlet 107, all of which are fluidly coupled via conduit. A fan 108 that generates a condenser airflow 126 (or pump 108 that generates a condenser liquid flow 126) cools refrigerant in the condenser. In this example, one or more low heat loads 120 are also in thermally conductive and/or convective contact with the cold plate 116. The optional solenoid valve 112 can be used when the expansion valve 114 is not configured to completely stop refrigerant flow when the TMS 800 is in an OFF state.

In some implementations of the CCRS 840, an oil is used for lubrication of the compressor 104 and the oil travels with the refrigerant in the closed-circuit portion of the TMS 800. The oil is removed from the refrigerant to be recirculated back to the compressor 104. The oil can be removed from the inlet 125 of the suction accumulator 124, within the suction accumulator 124, or elsewhere within the TMS 800. TMS 800 has a mechanism, e.g., a solenoid valve (not referenced) and an orifice 130, to return oil from the suction accumulator 124 (or a liquid separator), particularly, from the bottom of the suction accumulator 124 to the compressor 104. In addition, the CCRS 840 may include an oil separator. The oil separator is disposed in an oil return path from the compressor outlet 103 to the compressor inlet 101.

TMS 800 cools heat loads 118 and 120 (shown with the cold plate 116). The low heat load 120 is low heat flux that operates over long (or continuous) time intervals and are cooled by the CCRS 840, whereas the high heat load 118 is high heat flux that operates over short intervals of time relative to the operating interval of the low heat load 120.

When the low heat load 120 is applied, the TMS 800 is configured to have the CCRS 840 provide refrigeration to the low heat load 120. In this instance, the control system 999 produces signals to cause the solenoid valve 112 to be placed in an OFF state (i.e., closed). With the solenoid valve 112 closed, the CCRS 840 provides cooling duty to handle the low load 120.

In the closed-circuit refrigeration configuration, circulating refrigerant enters the compressor 104 as a saturated or superheated vapor and is compressed to a higher pressure at a higher temperature (a superheated vapor). This superheated vapor is at a temperature and pressure at which it can be condensed in the condenser 106 by either cooling water 126 or cooling air 126 flowing across a coil or tubes in the condenser 106. At the condenser 106, the circulating refrigerant loses heat and thus removes heat from the system, which removed heat is carried away by either the water or air (whichever may be the case) flowing over the coil or tubes, providing a condensed liquid refrigerant.

The condensed and sub-cooled liquid refrigerant is routed into the receiver 110, exits the receiver 110, and enters the expansion valve 114 (through the optional solenoid valve 112, if used). The refrigerant is enthalpically expanded in the expansion valve 114 and the high pressure sub-cooled liquid refrigerant turns into liquid-vapor mixture at a low pressure and temperature. The temperature of the liquid and vapor refrigerant mixture (evaporating temperature) is lower than the temperature of the low heat load 120. The mixture is routed through coils or tubes in the cold plate 116.

The heat from the low heat load 120 in contact with or proximate to the cold plate 116 partially or completely evaporates the liquid portion of the two-phase refrigerant mixture, and may superheat the mixture. The refrigerant leaves the cold plate 116 and enters the suction accumulator 124. The saturated or superheated vapor exits the outlet 127 of the suction accumulator 124 and enters the compressor 104. The cold plate 116 is where the circulating refrigerant absorbs and removes heat from the applied low heat loads, which heat is subsequently rejected in the condenser 106 and transferred to an ambient by water or air in the condenser 106. To complete the refrigeration cycle, the refrigerant vapor from the cold plate 116 is stored in the suction accumulator 124 and again a saturated vapor portion of the refrigerant in the suction accumulator 124 is routed from the compressor 104 back into the condenser 106.

On the other hand, when high heat load 118 is applied, a mechanism such as the control system 999 causes the TMS 800 to operate in both a closed and open-circuit configuration. The closed-circuit portion is similar to that described above, except that the cold plate 116 in this case operates within a threshold of a vapor quality, e.g., the cold plate 116 may operate with two phase mixture provided that the suction accumulator captures incidental non-evaporated liquid, and the compressor 104 receives saturated vapor from the suction accumulator 124 or the cold plate 116 may operate with a superheat.

When the TMS 800 operates with the open cycle, this causes the control system 999 to be configured to cause the back-pressure regulator 142 to be placed in an ON position, thus opening the back-pressure regulator 142 to permit the back-pressure regulator 142 to exhaust vapor through the exhaust line 140. The back-pressure regulator 142 maintains a back-pressure at an inlet to the back-pressure regulator 142, according to a set point pressure, while allowing the back-pressure regulator 142 to exhaust refrigerant vapor through the exhaust line 140.

The OCRS 850 operates like a thermal energy storage (TES) system, increasing cooling capacity of the TMS 800 when a pulsing heat load is activated, but without a duty cycle cooling penalty commonly encountered with TES systems. The cooling duty is executed without the concomitant penalty of conventional TES systems provided that the receiver 110 has enough refrigerant charge and the refrigerant flow rate flowing through the cold plate 116 matches the rate needed by the high loads. The back-pressure regulator 142 exhausts the refrigerant vapor less the refrigerant vapor recirculated by the compressor 104. The rate of exhaust of the refrigerant vapor through the exhaust line 140 is governed by the set point pressure used at the inlet to the back-pressure regulator 142.

When the high heat load 118 is no longer in use or its temperature is reduced, this occurrence is sensed by a sensor (not shown) and signal from the sensor (or otherwise, such as communicated directly by the high heat loads) are sent to the control system 999. The control system 999 is configured to partially or completely close the back-pressure regulator 142 by changing the set point pressure (or otherwise), partially or totally closing the exhaust line 140 to reduce or cut off exhaust refrigerant flow through the exhaust line 140. When the high heat load 118 reaches a desired temperature or is no longer being used, the back-pressure regulator 142 is placed in the OFF status and is thus closed, and CCRS 840 continues to operate as needed.

The provision of the CCRS 840 helps to reduce the amount of exhausted refrigerant (e.g., ammonia). Generally, the TMS 800 uses the compressor 104 to save ammonia and, in general, it may not be desirable to shut the compressor 104 off. For instance, the compressor 104 can help to keep a high pressure in the receiver 110 if a head pressure control valve is applied.

On the other hand, in some embodiments, the TMS 800 can be configured to operate in modes where the compressor 104 is turned off and the TMS 800 operates in open-circuit mode only (such as in fault conditions in the circuit or cooling requirements).

As further shown in FIG. 8 , in this example implementation of TMS 800, an optional sensor 134 is placed downstream of the outlet 123. In this example, the expansion valve 114 can be operated with the optional sensor 134 that controls the expansion valve 114 either directly or through control system 999. The cold plate 116 operates in two-phase (liquid/gas) for high heat load 118 and superheated regions for low heat load 120 with controlled superheat. The sensor-controlled expansion valve 114 and the sensor 134 provide a mechanism to measure and control superheat.

As further shown in FIG. 8 , an optional bypass valve 136 is positioned in an optional bypass 132 that is fluidly coupled between outlet 123 and inlet 125 (at one end) and between outlet 127 and junction 138 (at another end). The optional bypass valve 136 includes an inlet 129 and an outlet 131. In some aspects, optional bypass valve 136 can be opened, thereby allowing refrigerant fluid to bypass the suction accumulator 124. Further, in other alternative aspects of TMS 800, the suction accumulator 124 is optional (as is the optional bypass valve 136) and the outlet 123 can be in direct fluid communication with junction 138. The junction 138 is disposed in a flow circuit to connect the outlet 123 of the cold plate 116, the back-pressure regulator 142 and the compressor inlet 101 to the compressor 104. Compressor discharge is used to control pressure in the receiver 110 so that the pressure remains high enough to extend operation of the TMS 800, as the amount of liquid refrigerant in receiver 110 is consumed, reducing refrigerant pressure. If applied, the solenoid valve 112 can maintain a relatively constant pressure in the receiver 110 during the entire period of operation of the OCRS 850.

This example of TMS 800 also includes an optional recuperative heat exchanger 122. In this example implementation, the TMS 800 allows for a lower vapor quality at the outlet 123 because of the presence of the recuperative heat exchanger 122 that evaporates any remaining liquid prior to being fed to the inlet 101 of the compressor 104. In some implementations, the presence of the recuperative heat exchanger 122 can eliminate the need for the suction accumulator 124. The recuperative heat exchanger 122 is coupled in an input path 144 between the receiver 110 and the expansion valve 114 and in an output path 146 from vapor-side outlet 127 of the suction accumulator 124 (or just outlet 123 if no suction accumulator 124 is used) to junction 138. The expansion valve 114 can be operated with the sensor 134 (as discussed) that controls the expansion valve 114 either directly or indirectly via the control system 999. The solenoid valve 112 can maintain a relatively constant pressure in the receiver 110 during the entire period of operation of the TMS 800.

The recuperative heat exchanger 122 transfers heat energy from the refrigerant fluid emerging from suction accumulator 124 to refrigerant fluid upstream from the expansion valve 114. Inclusion of the recuperative heat exchanger 122 reduces mass flow rate demand and allows operation of cold plate 116 within the threshold of vapor quality. In some examples the recuperative heat exchanger 122 transfers heat energy from the refrigerant fluid emerging from cold plate 116, and the suction accumulator 124 is not needed. That is, the recuperative heat exchanger 122 obviates the need for the suction accumulator 124.

The discussion below regarding vapor quality presumes that the recuperative heat exchanger 122 is configured to generate a sufficient superheat and is used with the suction accumulator 124 to avoid presence of liquid at the compressor inlet 101. The vapor quality of the refrigerant fluid after passing through cold plate 116 can be controlled either directly or indirectly with respect to a vapor quality set point by the control system 999. The cold plate 116 may be configured to maintain exit vapor quality substantially below the critical vapor quality defined as “1.” Control of vapor quality is applicable to all disclosed embodiments.

Vapor quality is the ratio of mass of vapor to mass of liquid+vapor and is generally kept in a range of approximately 0.5 to almost 1.0; more specifically 0.6 to 0.95; more specifically 0.75 to 0.9 more specifically 0.8 to 0.9 or more specifically about 0.8 to 0.85. “Vapor quality” is thus defined as mass of vapor/total mass (vapor+liquid). In this sense, vapor quality cannot exceed “1” or be equal to a value less than “0.” In practice vapor quality may be expressed as “equilibrium thermodynamic quality” that is calculated as follows:

X=(h−h′)/(h″−h′),

where h is specific enthalpy, specific entropy or specific volume, h′ is of saturated liquid and ″ is of saturated vapor. In this case X can be mathematically below 0 or above 1, unless the calculation process is forced to operate differently. Either approach is acceptable.

For open circuit operation of TMS 800 using the recuperative heat exchanger 122, the refrigerant fluid emerging from cold plate 116 is transported through conduit to the recuperative heat exchanger 122. After passing through the recuperative heat exchanger 122, the refrigerant fluid is discharged as exhaust, via back-pressure regulator 142, through exhaust line 140 for TMS 800.

As discussed in the previous section, by adjusting the pressure pc of the refrigerant fluid, the temperature at which the liquid refrigerant phase undergoes vaporization within cold plate 116 can be controlled. Thus, in general, the temperature of heat loads 118 and 120 can be controlled by a device or component of TMS 800 that regulates the pressure of the refrigerant fluid within cold plate 116. TMS 800 operating parameters include the superheat and the vapor quality of the refrigerant fluid emerging from cold plate 116.

The vapor quality, which is a number from 0 to 1, represents the fraction of the refrigerant fluid that is in the vapor phase. Considering the high heat load 118 individually, because heat absorbed from high heat load 118 is used to drive a constant-temperature evaporation of liquid refrigerant to form refrigerant vapor in cold plate 116, it is generally important to ensure that, for a particular volume of refrigerant fluid propagating through cold plate 116, at least some of the refrigerant fluid remains in liquid form right up to the point at which the exit aperture of the cold plate 116 is reached to allow continued heat absorption from high heat load 118 without causing a temperature increase of the refrigerant fluid. If the fluid is fully converted to the vapor phase after propagating only partially through cold plate 116, further heat absorption by the (now vapor-phase) refrigerant fluid within cold plate 116 will lead to a temperature increase of the refrigerant fluid and high heat load 118.

On the other hand, liquid-phase refrigerant fluid that emerges from cold plate 116 represents unused heat-absorbing capacity, in that the liquid refrigerant fluid did not absorb sufficient heat from the high heat load 118 to undergo a phase change. To ensure that TMS 800 operates efficiently, the amount of unused heat-absorbing capacity should remain relatively small.

In addition, the boiling heat transfer coefficient that characterizes the effectiveness of heat transfer from the high heat load 118 to the refrigerant fluid is typically very sensitive to vapor quality. When the vapor quality increases from zero to a certain value, called a critical vapor quality, the heat transfer coefficient increases. When the vapor quality exceeds the critical vapor quality, the heat transfer coefficient is abruptly reduced to a very low value, causing dryout within cold plate 116. In this region of operation, the two-phase mixture behaves as superheated vapor.

In general, the critical vapor quality and heat transfer coefficient values vary widely for different refrigerant fluids, and heat and mass fluxes. For all such refrigerant fluids and operating conditions, the systems and methods disclosed herein control the vapor quality at the outlet of the evaporator such that the vapor quality approaches the threshold of the critical vapor quality.

To make maximum use of the heat-absorbing capacity of the two-phase refrigerant fluid mixture for high heat load 118, the vapor quality of the refrigerant fluid emerging from cold plate 116 should nominally be equal to the critical vapor quality. Accordingly, to both efficiently use the heat-absorbing capacity of the two-phase refrigerant fluid mixture and also ensure that the temperature of high heat load 118 remains approximately constant at the phase transition temperature of the refrigerant fluid in cold plate 116, the systems and methods disclosed herein are generally configured to adjust the vapor quality of the refrigerant fluid emerging from cold plate 116 to a value that is less than or equal to the critical vapor quality.

Another important operating consideration for TMS 800 is the mass flow rate of refrigerant fluid within the TMS 800. Cold plate 116 can be configured to provide minimal mass flow rate controlling maximal vapor quality, which is the critical vapor quality. By minimizing the mass flow rate of the refrigerant fluid according to the cooling requirements for high heat load 118, TMS 800 operates efficiently. Each reduction in the mass flow rate of the refrigerant fluid (while maintaining the same temperature set point value for high heat load 118) means that the charge of refrigerant fluid added to receiver 110 initially lasts longer, providing further operating time for TMS 800.

Within cold plate 116, the vapor quality of a given quantity of refrigerant fluid varies from the evaporator inlet (where vapor quality is lowest) to the evaporator outlet (where vapor quality is highest). Nonetheless, to realize the lowest possible mass flow rate of the refrigerant fluid within the TMS 800, the effective vapor quality of the refrigerant fluid within cold plate 116—even when accounting for variations that occur within cold plate 116—should match the critical vapor quality as closely as possible.

CCRS power demand and CCRS efficiency are optimal when the evaporating temperature is as high as possible and the condensing pressure is as low as possible. The condenser 106 and cold plate 116 dimensions can be reduced when the evaporating temperature is as low as possible and the condensing pressure is as high as possible.

To ensure that the OCRS 850 operates efficiently and the mass flow rate of the refrigerant fluid is relatively low, and at the same time the temperature of the high heat load 118 is maintained within a relatively small tolerance, TMS 800 adjusts the vapor quality of the refrigerant fluid emerging from cold plate 116 to a value such that an effective vapor quality within cold plate 116 matches, or nearly matches, the critical vapor quality. At the same time requirements for CCRS efficient operation would be taken into consideration as well. In addition, generally compressor 104 does not work well with liquids at inlet 101. Accordingly, operation of the compressor 104 as close as possible to the critical vapor quality is desirable.

In TMS 800, expansion valve 114 is generally configured to control the vapor quality of the refrigerant fluid emerging from cold plate 116. As an example, expansion valve 114 regulates the mass flow rate of the refrigerant fluid through the valve 114. In turn, for a given set of operating conditions (e.g., ambient temperature), initial pressure in the receiver 110, temperature set point value for high heat load 118, the vapor quality determines mass flow rate of the refrigerant fluid emerging from cold plate 116. Expansion valve 114 typically controls the vapor quality of the refrigerant fluid emerging from cold plate 116 in response to information about at least one thermodynamic quantity that is either directly or indirectly related to the vapor quality. In general, a wide variety of different measurement and control strategies can be implemented in TMS 800 to achieve the control objectives.

Referring now to FIG. 9 , an example implementation of a TMS 900 that includes an OCRS 950 fluidly coupled to a CCRS 940 with an ejector assist circuit 902 and one or more cold plates 116 is shown. In some aspects, the use of the ejector assist circuit 902 can assist in reducing a power requirement of the TMS 900. Items illustrated and referenced, but not mentioned in the discussion below, are discussed and referenced in FIG. 1 and FIG. 8 .

As shown in FIG. 9 , the TMS 900 includes the OCRS 950 integrated with the CCRS 940. The OCRS 950 is an ejector assisted open circuit refrigeration system. The CCRS 940 generally, includes ejector assist circuit 902 and a liquid separator 124 with a vapor section 916 and a liquid section 918 rather than a suction accumulator. The CCRS 940 provides cooling for low heat load 120 over long time intervals while the OCRS 950 provides cooling for high heat load 118 over short time intervals, as generally discussed above.

TMS 900 includes the receiver 110 that is configured to store sub-cooled liquid refrigerant, as discussed above, and may include an optional solenoid valve 112 and optional expansion valve 114. Both, either or neither of the optional solenoid valve 112 and the optional expansion valve 114 can be used (i.e., or not used) in example embodiments of the TMS 900.

The TMS 900 includes an ejector 930. The ejector 930 has a primary inlet (i.e., high pressure inlet) 901 that is coupled to the receiver 110 (either directly or through the optional solenoid valve 112 and/or expansion valve 114). Outlet 905 of the ejector 930 is coupled to the inlet 125 of the liquid separator 124 (through a cold plate 116 in this example, as explained more fully below). The ejector 930 also has a secondary inlet or low-pressure inlet 903 that is coupled to the liquid-side outlet 905 via the cold plate 116. The vapor-side outlet 127 of the liquid separator 124 is coupled to the junction 138 that is coupled to the back-pressure regulator 142. The back-pressure regulator 142 has an outlet (not referenced) that feeds exhaust line 140. The junction 138 is coupled to the inlet 101 of the compressor 104. The outlet 103 of the compressor 104 is coupled to the inlet 105 of the condenser 106.

In some aspects, ejector 930 includes a high-pressure motive nozzle or primary inlet 901, a suction or secondary inlet 903, a secondary nozzle that feeds a suction chamber, a mixing chamber for the primary flow of refrigerant and secondary flow of refrigerant to mix, and a diffuser. In one embodiment, the ejector 930 is passively controlled by built-in flow control. Also, optional flow control devices may be employed upstream of the ejector 930. Liquid refrigerant from the receiver 110 is the primary flow. In the motive nozzle, potential energy of the primary flow is converted into kinetic energy reducing the potential energy (the established static pressure) of the primary flow. The secondary flow from the outlet 123 of the cold plate 116 (or from the liquid outlet 915) has a pressure that is higher than the established static pressure in the suction chamber, and thus the secondary flow is entrained through the suction inlet 903 and the secondary nozzles internal to the ejector 930. The two streams (primary flow and secondary flow) mix together in the mixing chamber. In the diffuser section, the kinetic energy of the mixed streams is converted into potential energy elevating the pressure of the mixed flow liquid/vapor refrigerant that leaves the ejector 930 and is fed to the inlet 121 of the cold plate 116 (or liquid separator 124). In the context of open-circuit refrigeration systems, the use of the ejector 930 allows for recirculation of liquid refrigerant captured by the liquid separator 124 to increase the efficiency of the TMS 900. That is, by allowing for some recirculation of refrigerant, but without the need for a compressor or a condenser, this recirculation reduces the required amount of refrigerant needed for a given amount of cooling of ‘n-x’ high heat load 118 over given periods of operation.

The cold plate 116 is coupled between the ejector outlet 905 of the ejector 930 and the inlet 125 of the liquid separator 124 in this example ejector assist circuit 902. In this configuration, the cold plate 116 is coupled between the ejector outlet 905 and the liquid separator inlet 125. The recirculation rate in this example is equal to the vapor quality at the evaporator outlet 123. The expansion valve 114 or other control device can be built in the motive nozzle of the ejector 930 and provides active control of the thermodynamic parameters of refrigerant state at the outlet 123. By placing the cold plate 116 between the outlet 905 of the ejector 930 and the inlet 125 of the liquid separator 124, the necessity of having liquid refrigerant pass through the liquid separator 124 during the initial charging of the cold plate 116 with the liquid refrigerant is avoided. At the same time liquid trapped in the liquid separator 124 may be wasted after the TMS 900 shuts down.

In an alternative example of the ejector assist circuit 902 of the TMS 900, the cold plate 116 can be fluidly coupled such that the inlet 121 is connected to liquid outlet 915 of the liquid separator (e.g., through junction 914) and the outlet 123 fluidly coupled to the secondary inlet 903 of the ejector 930 (as shown in the dashed line representation of cold plate 116). For example, the cold plate 116 can be coupled (at outlet 123) to the secondary inlet 903 of the ejector 930 and (at inlet 121) to junction 904, to couple the cold plate 116 to the liquid-side outlet 915 of the liquid separator 124.

During open circuit operation in such a configuration, the ejector 930 again acts as a “pump,” to “pump” a secondary fluid flow, e.g., liquid/vapor from the cold plate 116 using energy of the primary refrigerant flow from the receiver 110. The cold plate 116 may be configured to maintain exit vapor quality below the critical vapor quality defined as “1.” However, the higher the exit vapor quality the better it is for operation of the ejector 930. Vapor quality is the ratio of mass of vapor to mass of liquid+vapor and is generally kept in a range of approximately 0.5 to almost 1.0; more specifically 0.6 to 0.95; more specifically 0.75 to 0.9 more specifically 0.8 to 0.9 or more specifically about 0.8 to 0.85. In such a configuration, refrigerant from receiver 110 enters into the primary inlet 901 of the ejector 930 and through the ejector assist circuit 902, meaning that refrigerant flows from the ejector 930 into liquid separator 124 and flow from the liquid separator 124 into the cold plate 116, which cools the heat loads 118 and/or 120. The refrigerant is returned to the ejector 930 and to the liquid separator 124, while a vapor fraction of the refrigerant is fed to the compressor 104 and to the condenser 106. The liquid separator 124 is used to insure only vapor exists at the input to the compressor 104. In this embodiment, an optional sensor 932 can be disposed at the outlet 123 of the cold plate 116 (shown in dashed line).

Further, in another alternative example of the ejector assist circuit 902 of the TMS 900, dual cold plates 116 can be installed in the ejector assist circuit 902 (i.e., with both the solid line and dashed line representations of the cold plate 116 in FIG. 9 included). In such an example, a cold plate 116 is coupled between outlet 905 of the ejector 930 and inlet 125 to the liquid separator 124 and another cold plate 116 is coupled between junction 904 and secondary inlet 903 to the ejector 930. Heat loads 118 and/or 120 are coupled to both cold plates 116. The cooling capacity of this configuration of TMS 900 may not be sensitive to recirculation rate (as compared to other example implementations), which may be beneficial when the heat loads may significantly reduce recirculation rate. An operating advantage of such a configuration is that by using dual cold plates 116, it is possible to run the cold plates 116 combining the features of the configurations mentioned above. Also, in this configuration, at least one cold plate 116 can operate with both high heat load 118 and low heat load 120 if those loads allow for operation in superheated regions.

Further, in another alternative example of the ejector assist circuit 902 of the TMS 900, the cold plate 116 can comprise a single cold plate 116 that is attached downstream from and upstream of the ejector 930.

In this example of the TMS 900, an optional evaporator circuit 920 is fluidly coupled at junction 914 to the ejector assist circuit 902. In this optional circuit 920, a conventional evaporator 910 (e.g., such as a simple single cold plate rather than a two-phase cold plate) is disposed within exhaust 904 with an inlet 931 and an outlet 933. The conventional evaporator 910 is in thermal conductive and/or convective contact with heat load 912. Optionally in the optional evaporator circuit 920 is an expansion valve 906 and sensor 908. The conventional evaporator 910 operates in a superheated region with controlled superheat by the expansion valve 906, which has a control port that is fed from sensor 908. The sensor 908 controls the expansion valve 906 and provides a mechanism to measure and control superheat.

If the optional expansion valve 906 and sensor 908 are not included with the optional evaporator circuit 920, then the conventional evaporator 910 shares the same control device 114, i.e., an expansion valve, as the cold plate 116 (or cold plates 116 in the case of dual cold plates 116). The cold plate(s) 116 operates in two-phase (liquid/gas) and conventional evaporator 910 operates in a superheated region with controlled superheat.

The CCRS 940 includes the receiver 110, optional solenoid valve 112, optional expansion valve 114, ejector 930, liquid separator 124, junction 138, compressor 104, and condenser 106, and includes the ejector assist circuit 902 comprised of the cold plate(s) 116 and junction 914, all coupled via conduit. The OCRS 950 portion of the TMS 900 includes the receiver 110, optional solenoid valve 112, optional expansion valve 114, ejector 930, liquid separator 124, junction 138, back-pressure regulator 142, exhaust line 140, and includes the ejector assist circuit 902 comprised of the cold plate(s) 116, all coupled via conduit. The TMS 900 is configured to cool the heat loads 118 and/or 120, as discussed above, that are coupled to and/or in thermal communication with the cold plate(s) 116.

In some embodiments, refrigerant flow through the TMS 900 during open-circuit operation is controlled in the OCRS 950 either solely by the ejector 930 and back-pressure regulator 142 or by those components aided by either one or all of the optional solenoid valve 112 and expansion valve 114, depending on requirements of the application, e.g., ranges of mass flow rates, cooling requirements, receiver capacity, ambient temperatures, heat load, etc.

While both expansion valve 114 and solenoid valve 112 may not typically be used, in some implementations, either or both would be used and would function as a flow control device(s) to control refrigerant flow into the primary inlet 901 of the ejector 930. In some embodiments, the expansion valve 114 can be integrated with the ejector 930. In various embodiments of the TMS 900, the optional expansion valve 114 may be required under some circumstances where there are or can be significant changes in, e.g., an ambient temperature, which might impose additional control requirements on the TMS 900.

The back-pressure regulator 142 has an outlet (not referenced) that is disposed at the exhaust line 140, and further has an inlet (not referenced) coupled via junction 138 to the vapor-side outlet 127 of the liquid separator 124. The back-pressure regulator 142 functions to control the vapor pressure upstream of the back-pressure regulator 142. In TMS 900, the back-pressure regulator 142 is a control device that controls the refrigerant fluid vapor pressure from the liquid separator 124 and indirectly controls evaporating pressure/temperature when the TMS 900 is operating in open circuit mode. In general, back-pressure regulator 142 can be implemented using a variety of different mechanical and electronic flow regulation devices, as mentioned above. The back-pressure regulator 142 regulates fluid pressure upstream from the regulator, i.e., regulates the pressure at the inlet to the back-pressure regulator 142 according to a set pressure point value.

Some loads require maintaining thermal contact between the high heat load 118 and cold plate 116 with the refrigerant being in the two-phase region (of a phase diagram for the refrigerant) and, therefore, the flow control device 114 maintains a proper vapor quality at the evaporator outlet. Alternatively, a sensor communicating with control system 999 may monitor pressure in the receiver 110, as well as a pressure differential across the expansion valve 114, a pressure drop across the cold plate 116, a liquid level in the liquid separator 124, and power input into electrically actuated heat loads, or a combination of the above.

The CCRS 940 operates like CCRS 840 as described with reference to FIG. 8 , except that refrigerant from receiver 110 enters into the primary inlet 901 of the ejector 930 and through the ejector assist circuit 902, meaning that refrigerant flows from the ejector 930 into liquid separator 124 (and possibly through cold plate 116 beforehand) and flow from the liquid separator 124 into the cold plate 116 (if positioned as shown in dashed line representation or in a dual cold plate configuration), which cools heat loads 118 and/or 120. The refrigerant is returned to the ejector 930 and to the liquid separator 124, while a vapor fraction of the refrigerant is fed to the compressor 104 and to the condenser 106, as discussed above. The liquid separator 124 is used to insure only vapor exists at the input to the compressor 104.

The OCRS 950 with the ejector 930 operates as follows. The liquid refrigerant from the receiver 110 (primary flow) is fed to the primary inlet 901 of the ejector 930 and expands at a constant entropy in the ejector 930 (in ideal case; in reality the nozzle is characterized by the isentropic efficiency of the ejector) and turns into a two-phase (gas/liquid) state. The refrigerant in the two-phase state from the ejector 930 enters the liquid separator 124 (and possibly through cold plate 116 beforehand), at inlet port 125 with only or substantially only liquid exiting the liquid separator 124 at the liquid-side outlet 915 and only or substantially only vapor exiting the separator 124 at vapor-side outlet 127. The liquid stream exiting at outlet 915 enters into a liquid/vapor stream that enters the cold plate 116 (if positioned as shown in dashed line representation or in a dual cold plate configuration).

The cold plate 116 provides cooling duty and discharges the refrigerant in a two-phase state at relatively low exit vapor quality (low fraction of vapor to liquid, e.g., generally below 0.5) into the secondary inlet 903 of the ejector 930. The ejector 930 entrains the refrigerant flow exiting the cold plate 116 and combines it with the primary flow from the receiver 110. Vapor exits from the vapor-side outlet 127 of the liquid separator 124 and is exhausted by the exhaust line 140. The back-pressure regulator 142 regulates the pressure upstream of the back-pressure regulator 142 so as to maintain upstream refrigerant fluid pressure in TMS 900.

Refrigerant from the receiver 110 is directed into the ejector 930 (optionally through valve 112 and the expansion valve 114) and expands at a constant entropy in the ejector 930 (in an ideal case; in reality the nozzle is characterized by the ejector isentropic efficiency), and turns into a two-phase (gas/liquid) state. The refrigerant in the two-phase state enters the cold plate 116 that cools the heat loads 118 and/or 120 and discharges the refrigerant in a two-phase state at an exit vapor quality (fraction of vapor to liquid) below a unit vapor quality (“1”). The discharged refrigerant is fed to the inlet 125 of the liquid separator 124, where the liquid separator 124 separates the discharge refrigerant with only or substantially only liquid exiting the liquid separator at the liquid-side outlet 915 and only or substantially only vapor exiting the separator 124 at the vapor-side outlet 127. The vapor-side may contain some liquid droplets since the liquid separator 124 has a separation efficiency below a “unit” separation. The liquid stream exiting at outlet 915 enters the suction or secondary inlet 903 of the ejector 930. The ejector 930 entrains the refrigerant flow exiting the expansion valve by the refrigerant from the receiver 110.

In closed circuit operation, back-pressure regulator 142 is turned off and vapor from the liquid separator 124 is fed to the compressor 104 and condenser 106, as generally discussed above. In open circuit operation, back-pressure regulator 142 is turned on and a portion of the vapor is exhausted through exhaust line 140, as generally discussed above.

Referring now to FIG. 10 , an example implementation of a TMS 100 that includes an OCRS 1050 fluidly coupled to a CCRS 1040 with a pump assist circuit 1002 and one or more cold plates 116 is shown. In some aspects, the use of the pump assist circuit 902 can assist in reducing a power requirement of the TMS 1000. Items illustrated and referenced, but not mentioned in the discussion below, are discussed and referenced in FIG. 1 , FIG. 8 , and FIG. 9 . The CCRS 1040 provides cooling for low heat load 120 over long time intervals while the OCRS 1040 provides cooling for high heat load 118 over short time intervals, as generally discussed above.

TMS 1000 includes the receiver 110 that is configured to store liquid refrigerant, i.e., subcooled liquid refrigerant, optional solenoid control valve 112, optional expansion valve 114, and a junction 1004 that has first and second ports configured as inlets and a third port configured as an outlet. TMS 1000 also includes cold plate 116, liquid separator 124, a pump 1008 having inlet 1003 and outlet 1005, back-pressure regulator 142, and exhaust line 140. TMS 1000 also includes compressor 104 and the condenser 106 having the outlet 107 coupled to the inlet 109 of receiver 110. The TMS includes pump assist circuit 1002 having the junction 1004, one or more cold plates 116, the liquid separator 124, and the pump 1008.

The junction 1004 has the first port coupled to the receiver 110 (e.g., through optional valve 112 and expansion valve 114), the second port as an inlet coupled to the outlet 123 of the cold plate 116 (shown in solid line), and a third port as the outlet coupled to the inlet 125 of the liquid separator 124 (through an optional second cold plate 116).

The liquid separator 124 has the inlet 125, the vapor-side outlet 127 and liquid-side outlet 915. The vapor-side outlet 127 of the liquid separator 124 is coupled via junction 138 to inlet 102 of the compressor 104 that controls a vapor pressure in the cold plate 116 and feeds vapor to the condenser 106. The vapor-side outlet 127 is coupled to one port of the junction 138 that feeds compressor 104 and the back-pressure regulator 142. The back-pressure regulator 142 has an outlet that feeds exhaust line 140. The liquid-side outlet 915 of the liquid separator 124 is coupled to inlet 1003 of the pump 1008 (as shown in this example).

The liquid separator 124 and pump 1008 can be arranged in several example configurations. For example, the liquid separator 124 (e.g., implemented as a flash drum) can have the pump 1008 located distal from the liquid-side port 915 as shown in FIG. 10 . This configuration potentially presents the possibility of cavitation. To minimize the possibility of cavitation, the pump 1008 can be located distal from the liquid-side outlet port 915, but the height at which the inlet 125 is located is higher than conventional. This would result in an increase in liquid pressure at the liquid-side outlet 915 of the liquid separator 124 and concomitant therewith an increase in liquid pressure at the inlet 1003 of the pump 1008. Increasing the pressure at the inlet 1003 to the pump 1008 should minimize possibility of cavitation. Another strategy is to locate the pump 1008 proximate to or indeed, inside of the liquid-side outlet 915. In addition, the height at which the inlet 125 is located can be adjusted increase liquid pressure at the inlet 1003 of the pump 1008 further minimizing the possibility of cavitation. Another alternative strategy that can be used for any of the configurations involves the use of a sensor that produces a signal that is a measure of the height of a column of liquid in the liquid separator 124. The signal is sent to the control system 999 that will be used to start the pump 1008, once a sufficient height of liquid is contained by the liquid separator 124.

Various types of pumps can be used for pump 1008. Exemplary types include gear, centrifugal, rotary vane types. When choosing a pump, the pump should be capable to withstand the expected fluid flows, including criteria such as temperature ranges for the fluids, and materials of the pump should be compatible with the properties of the fluid. A subcooled refrigerant can be provided at the pump outlet 1005 to avoid cavitation. To do that a certain liquid level in the liquid separator 124 may provide hydrostatic pressure corresponding to that sub-cooling.

The junction 1004 can positioned in several example positions in the pump assist circuit 1002. For example, one of the inlets and the outlet can be interposed between solenoid valve 112 and expansion valve 114, with its other inlet coupled to the outlet 123 of the cold plate 116 (shown in solid line). As another example, one of the inlets and the outlet interposed between the outlet 119 of the expansion valve 114 and inlet 121 to the cold plate 116 (shown dashed line) or inlet 125 to liquid separator 124, with its other inlet coupled to the outlet 123 of the cold plate 116 (shown in solid line). As another example, if both of the optional solenoid control valve 112 and optional expansion valve 114 are not included, then all of the locations for the junction 1004 are, in essence, the same provided that there are no other intervening functional devices between the outlet 111 of the receiver 110 and the inlet of the junction 1004.

In TMS 1000, refrigerant liquid from the liquid-side outlet 905 of the liquid separator 124 is fed to pump inlet 1003 and is pumped from the pump 1008 into the inlet 121 of the cold plate 116. Refrigerant exiting from the outlet 123 is fed along with the primary refrigerant flow from the expansion valve 114 back to the liquid separator 124 (through another cold plate 116 as an option as shown). These liquid refrigerant streams from the receiver 110 and the pump 1008 are mixed downstream from the expansion device 114. Heat loads 118 and/or 120 are in thermal conductive and/or convective contact with or in proximity to the cold plate 116 (or dual cold plates 116 as shown). The cold plate 116 is configured to extract heat from the heat loads 118 and/or 120 and to control the vapor quality at the outlet 123 of the cold plate 116.

The cold plate 116 is coupled between the pump outlet 1005 of the pump 1008 and the junction 1004 in this example pump assist circuit 1002. In an alternative example of TMS 1000, there can be dual cold plates 116 (with one shown in solid line in FIG. 10 and another cold plate 116 shown in dashed line as an option). Thus, in this example modification, a first cold plate 116 is coupled between the outlet of the junction 1004 and the inlet 125 of the liquid separator 124 (as shown in dashed line) and there is a second cold plate 116 having an inlet 121 that is coupled to the outlet 1005 of the pump 1008 and having an outlet 123 coupled to a second inlet of the junction 1004 (this cold plate 116 being shown in solid line). The liquid separator 124 has the inlet 125, the vapor-side outlet 127 and liquid-side outlet 915. Heat loads 118 and/or 120 are in thermal conductive and/or convective contact or in proximity to both cold plates 116 in this modified example. An operating advantage of this modified example is that by placing cold plates 116 at both the outlet and the second inlet of the junction 1004, it is possible to combine heat loads 118 and 120 at both cold plates 116, which requires operation in a two-phase region and which allows operation with superheat.

Further, in another alternative example of the pump assist circuit 1002 of the TMS 1000, the cold plate 116 can comprise a single cold plate 116 that is attached downstream from and upstream of the junction 1004. For example, as a single cold plate 116, the cold plate 116 has a first inlet 121 that is coupled to the outlet of the junction 1004 and a first outlet 123 that is coupled to the inlet 125 of the liquid separator 124. The cold plate 116 also has a second inlet 121 that is coupled to the outlet 1005 of the pump 1008 and has a second outlet 123 that is coupled to the inlet of the junction 1004. The liquid-side outlet 915 of the liquid separator 124 is coupled to the inlet 1003 of the pump 1008.

In this example implementation, optional evaporator circuit 920 is fluidly coupled to the liquid separator 124. For example, as shown, in this optional configuration, the liquid separator 124 is configured to have a second, liquid-side outlet 1001 in addition to the inlet 125, the vapor-side outlet 127, and the liquid-side outlet 915. Alternatively, such a function could be provided with another junction (not shown). The second outlet 1001 diverts a portion of the liquid exiting the liquid separator 124 into the conventional evaporator 910 that is in thermal contact with heat load 912. The conventional evaporator 910 extracts heat from the heat load 912 and exhausts vapor from exhaust line 904.

Exhaust lines 140 and 904 can be combined or can be separated. As shown in the optional evaporator circuit 920, in the case of exhaust 904 not being combined with exhaust 140, another back-pressure regulator 1006 can be placed in the exhaust 904.

The evaporator circuit 920 can cool heat loads in two-phase and superheated regions. The conventional evaporator 910 can be fed a portion of the liquid refrigerant and operate in superheated region without the need for active superheat control. For example, optionally in the optional evaporator circuit 920 is expansion valve 906 and sensor 908. The conventional evaporator 910 operates in a superheated region with controlled superheat by the expansion valve 906, which has a control port that is fed from sensor 908. The sensor 908 controls the expansion valve 906 and provides a mechanism to measure and control superheat.

The sensor 908 disposed proximate to the outlet 933 of the conventional evaporator 910 provides a measurement of superheat, and indirectly, vapor quality. For example, sensor 908 is a combination of temperature and pressure sensors that measures the refrigerant fluid superheat downstream from the heat load 912 and transmits the measurements to the control system 999. The control system 999 adjusts the expansion valve 906 based on the measured superheat relative to a superheat set point value. By doing so, control system 999 indirectly adjusts the vapor quality of the refrigerant fluid emerging from conventional evaporator 910.

In closed-circuit operation, the CCRS 1040 operates as follows. The back-pressure regulator 142 is placed in an OFF position. The liquid refrigerant from the receiver 110 is fed to the expansion valve 114 (if used) and expands at a constant enthalpy in the expansion valve 114 turning into a two-phase (gas/liquid) mixture. This two-phase liquid/vapor refrigerant is fed to the inlet 125 of the liquid separator 124 (or in the case of dual cold plates, the cold plate 116 shown in dashed line), where the liquid separator 124 separates the discharge refrigerant with only or substantially only liquid exiting the liquid separator 124 at the liquid-side outlet 915 (or both outlet 915 and outlet 1001 in the case of the optional evaporator circuit 920) and only or substantially only vapor exiting the liquid separator 124 at vapor-side outlet 127. The liquid stream exiting at liquid-side outlet 915 enters and is pumped by the pump 1008 into the cold plate 116 that provides cooling duty and discharges the refrigerant in a two-phase state at a relatively high exit vapor quality (fraction of vapor to liquid). The discharged refrigerant is fed to the junction 1004. Vapor from the vapor-side 127 of the liquid separator 124 is fed to the compressor 104, on to the condenser 106, and back into the receiver 110 for closed circuit operation.

On the other hand, when high heat load 118 is applied, a mechanism such as the control system 999 causes the TMS 1000 to operate in both a closed and open cycle configuration. The closed cycle portion would be similar to that described. The OCRS 1050 has the control system 999 configured to cause the back-pressure regulator 142 to be placed in an ON position, opening the back-pressure regulator 142 to permit the back-pressure regulator 142 to exhaust vapor through the exhaust line 140. The back-pressure regulator 142 maintains a back-pressure at an inlet to the back-pressure regulator 142, according to a set point pressure, while allowing the back-pressure regulator 142 to exhaust refrigerant vapor to the exhaust line 140.

In OCRS 1040, the pump 1008 can operate across a reduced pressure differential (pressure difference between inlet 1003 and outlet 1005 of the pump 1008). In the context of open circuit refrigeration systems, the use of the pump 1008 allows for recirculation of liquid refrigerant from the liquid separator 124 to enable operation at reduced vapor quality at the cold plate 116 outlet, avoiding the discharge of remaining liquid out of the TMS 1000 at less than the separation efficiency of the liquid separator 124 allows. This recirculation reduces the required amount of refrigerant needed for a given amount of cooling over a given period of operation. The configuration above reduces the vapor quality at the cold plate 116 inlet 121 and thus may improve refrigerant distribution (of the two-phase mixture) in the cold plate 116.

A variety of different refrigerant fluids can be used in TMS 100, 800, 900, or 1000. Depending on the application for both open circuit refrigeration system operation and closed-circuit refrigeration system operation, emissions regulations and operating environments may limit the types of refrigerant fluids that can be used.

For example, in certain embodiments, the refrigerant fluid can be ammonia having very large latent heat; after passing through the cooling circuit, the ammonia refrigerant vapor in the open circuit operation can be disposed of by incineration, by chemical treatment (i.e., neutralization), and/or by direct venting to the atmosphere. In certain embodiments, the refrigerant fluid can be an ammonia-based mixture that includes ammonia and one or more other substances. For example, mixtures can include one or more additives that facilitate ammonia absorption or ammonia burning.

More generally, any fluid can be used as a refrigerant in the open circuit refrigeration systems disclosed herein, provided that the fluid is suitable for cooling heat loads 118 and/or 120 (e.g., the fluid boils at an appropriate temperature) and, in embodiments where the refrigerant fluid is exhausted directly to the environment, regulations and other safety and operating considerations do not inhibit such discharge.

One example of refrigerant is ammonia. Ammonia under standard conditions of pressure and temperature is in a liquid or two-phase state. Thus, the receiver 110 typically will store ammonia at a saturated pressure corresponding to the surrounding temperature. The pressure in the receiver 110 storing ammonia will change during operation. The use of the first expansion valve 114 can stabilize pressure in the receiver 110 during operation, by adjusting the first expansion valve 114 (e.g., automatically or by control system 999) based on a measurement of the evaporation pressure (p_(e)) of the refrigerant fluid and/or a measurement of the evaporation temperature of the refrigerant fluid.

Control system 999 can adjust expansion valve 114 based on measurements of one or more of the following TMS parameter values: the pressure drop (p_(r)−p_(e)) across expansion device 18, the pressure drop across cold plate 116, the refrigerant fluid pressure in receiver 110 (p_(r)), the vapor quality of the refrigerant fluid emerging from cold plate 116 (or at another location in the TMS), the superheat value of the refrigerant fluid in the TMS, the evaporation pressure (p_(e)) of the refrigerant fluid, and the evaporation temperature of the refrigerant fluid.

To adjust first expansion valve 114 based on a particular value of a measured TMS parameter value, control system 999 compares the measured value to a set point value (or threshold value) for the TMS parameter, as will be discussed below.

A variety of different refrigerant fluids can be used in any of the open circuit configurations. For open circuit refrigeration systems in general, emissions regulations and operating environments may limit the types of refrigerant fluids that can be used. For example, in certain embodiments, the refrigerant fluid can be ammonia having very large latent heat; after passing through the cooling circuit, vaporized ammonia that is captured at the vapor port of the liquid separator can be disposed of by incineration, by chemical treatment (i.e., neutralization), and/or by direct venting to the atmosphere. Any liquid captured in the liquid separator is recycled back into the OCRS (either directly or indirectly).

Since liquid refrigerant temperature is sensitive to ambient temperature, the density of liquid refrigerant changes even though the pressure in the receiver 110 remains the same. Also, the liquid refrigerant temperature impacts the vapor quality at the evaporator inlet. Therefore, the refrigerant mass and volume flow rates change and the control devices can be used.

Another alternative strategy that can be used for any of the configurations depicted involves the use of a heat exchanger. The heat exchanger is an evaporator, which brings in thermal contact two refrigerant streams. In the above systems, a first of the streams is the liquid stream leaving the liquid separator 124. A second stream is the liquid refrigerant expanded to a pressure lower than the evaporator pressure in the cold plate 116 and evaporating the related evaporating temperature lower than the liquid temperature at the liquid separator exit. Thus, the liquid from the liquid-side outlet is subcooled rejecting thermal energy to the second side of the heat exchanger. The second side absorbs the rejected thermal energy due to evaporating and superheating of the second refrigerant stream.

Various combinations of the sensors can be used to measure thermodynamic properties of the TMS that are used to adjust the control devices or pumps discussed above and which signals are processed by the control system 999. Connections (wired or wireless) are provided between each of the sensors and control system 999. In many embodiments, the TMS includes only certain combinations of the sensors (e.g., one, two, three, or four of the sensors) to provide suitable control signals for the control devices.

Cooling of high heat flux that are also highly temperature sensitive can present a number of challenges. On one hand, such loads generate significant quantities of heat that is extracted during cooling. In conventional closed-cycle refrigeration systems, cooling high heat flux typically involves circulating refrigerant fluid at a relatively high mass flow rate. However, closed-cycle system components that are used for refrigerant fluid circulation—including compressors and condensers—are typically heavy and consume significant power. As a result, many closed-cycle systems are not well suited for deployment in mobile platforms—such as on small vehicles—where size and weight constraints may make the use of large compressors and condensers impractical.

Temperature sensitive loads such as electronic components and devices may require temperature regulation within a relatively narrow range of operating temperatures. Maintaining the temperature of such a load to within a small tolerance of a temperature set point can be challenging when a single-phase refrigerant fluid is used for heat extraction, since the refrigerant fluid itself will increase in temperature as heat is absorbed from the load.

Directed energy systems that are mounted to mobile vehicles such as trucks may present many of the foregoing operating challenges, as such systems may include high heat flux, temperature sensitive components that require precise cooling during operation in a relatively short time. The TMS implementations disclosed herein, while generally applicable to the cooling of a wide variety of heat loads, is particularly well suited for operation with such directed energy systems.

In particular, the thermal management systems and methods disclosed herein include a number of features that reduce both overall size and weight relative to conventional refrigeration systems, and still extract excess heat energy from both high heat flux, highly temperature sensitive components and relatively temperature insensitive components, to accurately match temperature set points for the components. At the same time the disclosed thermal management systems require no significant power to sustain their operation. Whereas certain conventional refrigeration systems used closed-circuit refrigerant flow paths, the systems and methods disclosed herein use open-cycle refrigerant flow paths. Depending upon the nature of the refrigerant fluid, exhaust refrigerant fluid may be incinerated as fuel, chemically treated, and/or simply discharged at the end of the flow path.

In the thermal management systems disclosed herein, a receiver 110 is initially charged with a refrigerant fluid that is in a liquid state. During operation of the system, the refrigerant fluid is transported from the receiver 110 through an open-cycle refrigerant flow path, and then discharged from an exhaust line. Effectively, the pressure of the refrigerant fluid in the receiver 110 functions as the driving force for mass transport of the refrigerant fluid through the system, as the thermal management systems do not use a pump or other mechanical device to drive refrigerant fluid flow from the receiver 110.

Typically, at the beginning of system operation, the refrigerant pressure in the receiver 110 is sufficient to drive refrigerant fluid at a mass flow rate sufficient to provide adequate cooling capacity for one or more loads connected to the system. As operation continues, however, the refrigerant pressure in the receiver 110 falls, owing to the continued transport of refrigerant fluid out of the receiver 110. Consequently, the maximum mass flow rate of refrigerant fluid that can be achieved falls. If operation continues for a sufficiently long period of time, the refrigerant pressure in the receiver 110 may no longer be adequate to support a desired cooling capacity for the connected loads, even if some refrigerant fluid remains in the receiver 110.

Moreover, the refrigerant pressure in the receiver 110 varies according to temperature. When the temperature of the environment within which the thermal management system is operated is relatively lower (such that the refrigerant fluid within the receiver 110 is also relatively lower), the refrigerant pressure in the receiver 110 is also lower and, as a result, the refrigerant fluid in the receiver 110 supports a relatively lower maximum mass flow rate of refrigerant fluid through the thermal management system. Even at the beginning of system operation, if the refrigerant fluid in the receiver 110 is at low enough temperature, the refrigerant pressure may be inadequate to support a refrigerant fluid mass flow rate that achieves a particular necessary or desirable cooling capacity for one or more heat loads connected to the thermal management system.

Typically, as the refrigerant pressure in the receiver 110 falls during operation of the thermal management system, a relatively complex series of control actions involving at least two control devices is implemented on an ongoing basis to ensure that the thermal management system continues to provide adequate cooling capacity for one or more connected heat loads. These control actions can involve, for example, adjusting the vapor quality of the refrigerant fluid and the temperature of one or more of the heat loads. To maintain these parameter values within a desired range even as the refrigerant pressure in the receiver 110 changes, the control devices can dynamically adjust refrigerant fluid flow rates in different system components.

To ensure that the thermal management systems disclosed herein can provide adequate cooling capacity even during start-up at relatively low temperatures, and to reduce the amount of unused refrigerant fluid that remains in the receiver 110 when operation of the thermal management systems extends to completion, the thermal management systems disclosed herein can optionally include one or more gas receivers 10 that are charged with one or more inert gases. The gas receiver(s) 10 is/are connected to the receiver 110, and gas from the gas receiver(s) 10 is transported into the receiver 110 to increase the total pressure in the receiver 110. Because the total pressure effectively functions as the driving force for refrigerant fluid transport through the thermal management system, the use of one or more gas receivers 10 can extend the operating time of the thermal management systems disclosed herein.

Referring to FIG. 11 , the example control system 999 includes a processor 1102, memory 1104, storage 1106, and I/O interfaces 1108, all of which are connected/coupled together via a bus 1110. Control system 999 can be used with any of the embodiments discussed herein, e.g., any of FIGS. 1 and 8-10 .

Control system 999 can generally, and optionally, include any one or more of a processor (or multiple processors), a memory, a storage device, and input/output device. Some or all of these components can be interconnected using a system bus. The processor is capable of processing instructions for execution. In some embodiments, the processor is a single-threaded processor. In certain embodiments, the processor is a multi-threaded processor. Typically, the processor is capable of processing instructions stored in the memory or on the storage device to display graphical information for a user interface on the input/output device, and to execute the various monitoring and control functions discussed above. Suitable processors for the systems disclosed herein include both general and special purpose microprocessors, and the sole processor or one of multiple processors of any kind of computer or computing device.

The memory stores information within the system, and can be a computer-readable medium, such as a volatile or non-volatile memory. The storage device can be capable of providing mass storage for the control system 999. In general, the storage device can include any non-transitory tangible media configured to store computer readable instructions. For example, the storage device can include a computer-readable medium and associated components, including: magnetic disks, such as internal hard disks and removable disks; magneto-optical disks; and optical disks. Storage devices suitable for tangibly embodying computer program instructions and data include all forms of non-volatile memory including by way of example, semiconductor memory devices, such as EPROM, EEPROM, and flash memory devices; magnetic disks such as internal hard disks and removable disks; magneto-optical disks; and CD-ROM and DVD-ROM disks. Processors and memory units of the systems disclosed herein can be supplemented by, or incorporated in, ASICs (application-specific integrated circuits).

The input/output device provides input/output operations for control system 999, and can include a keyboard and/or pointing device. In some embodiments, the input/output device includes a display unit for displaying graphical user interfaces and system related information.

The features described herein, including components for performing various measurement, monitoring, control, and communication functions, can be implemented in digital electronic circuitry, or in computer hardware, firmware, or in combinations of them. Methods steps can be implemented in a computer program product tangibly embodied in an information carrier, e.g., in a machine-readable storage device, for execution by a programmable processor (e.g., of control system 999), and features can be performed by a programmable processor executing such a program of instructions to perform any of the steps and functions described above. Computer programs suitable for execution by one or more system processors include a set of instructions that can be used directly or indirectly to cause a processor or other computing device executing the instructions to perform certain activities, including the various steps discussed above.

Computer programs suitable for use with the systems and methods disclosed herein can be written in any form of programming language, including compiled or interpreted languages, and can be deployed in any form, including as stand-alone programs or as modules, components, subroutines, or other units suitable for use in a computing environment.

Any two of the optional devices, as pressure sensors, upstream and downstream from a control device, can be configured to measure information about a pressure differential p_(r)-p_(e) across the respective control device and to transmit electronic signals corresponding to the measured pressure from which a pressure difference information can be generated by the control system 999. Other sensors such as flow sensors and temperature sensors can be used as well. In certain embodiments, sensors can be replaced by a single pressure differential sensor, a first end of which is connected adjacent to an inlet and a second end of which is connected adjacent to an outlet of a device to which differential pressure is to be measured, such as the evaporator. The pressure differential sensor measures and transmits information about the refrigerant fluid pressure drop across the device, e.g., the cold plate 116.

Temperature sensors can be positioned adjacent to an inlet or an outlet of e.g., the cold plate 116 or between the inlet and the outlet. Such a temperature sensor measures temperature information for the refrigerant fluid within cold plate 116 (which represents the evaporating temperature) and transmits an electronic signal corresponding to the measured information. A temperature sensor can be attached to any of heat loads 118 and/or 120, and/or 912 to measure temperature information for the load and transmits an electronic signal corresponding to the measured information. An optional temperature sensor can be adjacent to the outlet of cold plate 116 that measures and transmits information about the temperature of the refrigerant fluid as it emerges from cold plate 116.

In certain embodiments, the thermal management systems disclosed herein are configured to determine superheat information for the refrigerant fluid based on temperature and pressure information for the refrigerant fluid measured by any of the sensors disclosed herein. The superheat of the refrigerant vapor refers to the difference between the temperature of the refrigerant fluid vapor at a measurement point in a TMS and the saturated vapor temperature of the refrigerant fluid defined by the refrigerant pressure at the measurement point in a TMS.

To determine the superheat associated with the refrigerant fluid, the system control system 999 (as described) receives information about the refrigerant fluid vapor pressure after emerging from a heat exchanger downstream from cold plate 116, and uses calibration information, a lookup table, a mathematical relationship, or other information to determine the saturated vapor temperature for the refrigerant fluid from the pressure information. The control system 999 also receives information about the actual temperature of the refrigerant fluid, and then calculates the superheat associated with the refrigerant fluid as the difference between the actual temperature of the refrigerant fluid and the saturated vapor temperature for the refrigerant fluid.

The foregoing temperature sensors can be implemented in a variety of ways in a TMS. As one example, thermocouples and thermistors can function as temperature sensors in a TMS. Examples of suitable commercially available temperature sensors for use in a TMS include, but are not limited to, the 88000 series thermocouple surface probes (available from OMEGA Engineering Inc., Norwalk, Conn.).

A TMS can include a vapor quality sensor that measures vapor quality of the refrigerant fluid emerging from cold plate 116. Typically, such a sensor is implemented as a capacitive sensor that measures a difference in capacitance between the liquid and vapor phases of the refrigerant fluid. The capacitance information can be used to directly determine the vapor quality of the refrigerant fluid (e.g., by system control system 999). Alternatively, sensors can determine the vapor quality directly based on the differential capacitance measurements and transmit an electronic signal that includes information about the refrigerant fluid vapor quality. Examples of commercially available vapor quality sensors that can be used in a TMS include, but are not limited to, HBX sensors (available from HB Products, Hasselager, Denmark).

It should generally understood that the systems disclosed herein can include a variety of combinations of the various sensors described above, and control system 999 can receive measurement information periodically or aperiodically from any of the various sensors. Moreover, it should be understood any of the sensors described can operate autonomously, measuring information and transmitting the information to control system 999 (or directly to the first and/or second control device) or, alternatively, any of the sensors described above can measure information when activated by control system 999 via a suitable control signal, and measure and transmit information to control system 999 in response to the activating control signal.

To adjust a control device on a particular value of a measured system parameter value, control system 999 compares the measured value to a set point value (or threshold value) for the system parameter. Certain set point values represent a maximum allowable value of a system parameter, and if the measured value is equal to the set point value (or differs from the set point value by 10% or less (e.g., 5% or less, 3% or less, 1% or less) of the set point value), control system 999 adjusts a respective control device to modify the operating state of a TMS. Certain set point values represent a minimum allowable value of a system parameter and, if the measured value is equal to the set point value (or differs from the set point value by 10% or less (e.g., 5% or less, 3% or less, 1% or less) of the set point value), control system 999 adjusts the respective control device to modify the operating state of a TMS, and increase the system parameter value. The control system 999 executes algorithms that use the measured sensor value(s) to provide signals that cause the various control devices to adjust refrigerant flow rates, etc.

Some set point values represent “target” values of system parameters. For such system parameters, if the measured parameter value differs from the set point value by 1% or more (e.g., 3% or more, 5% or more, 10% or more, 20% or more), control system 999 adjusts the respective control device to adjust the operating state of the system, so that the system parameter value more closely matches the set point value.

Optional pressure sensors are configured to measure information about the pressure differential p_(r)-p_(e) across a control device and to transmit an electronic signal corresponding to the measured pressure difference information. Two sensors can effectively measure p_(r), p_(e). In certain embodiments two sensors can be replaced by a single pressure differential sensor. Where a pressure differential sensor is used, a first end of the sensor is connected upstream of a first control device and a second end of the sensor is connected downstream from first control device.

A TMS also includes optional pressure sensors positioned at the inlet and outlet, respectively, of cold plate 116. A sensor measures and transmits information about the refrigerant fluid pressure upstream from cold plate 116, and a sensor measure and transmit information about the refrigerant fluid pressure downstream from cold plate 116. This information can be used (e.g., by a system controller) to calculate the refrigerant fluid pressure drop across cold plate 116. As above, in certain embodiments, sensors can be replaced by a single pressure differential sensor to measure and transmit the refrigerant fluid pressure drop across cold plate 116.

To measure the evaporating pressure (p_(e)) a sensor can be optionally positioned between the inlet and outlet of cold plate 116, i.e., internal to cold plate 116. In such a configuration, the sensor can provide a direct measurement of the evaporating pressure.

To measure refrigerant fluid pressure at other locations within a TMS, sensor can also optionally be positioned, for example, in-line along a conduit. Pressure sensors at each of these locations can be used to provide information about the refrigerant fluid pressure downstream from cold plate 116, or the pressure drop across cold plate 116.

It should be appreciated that, in the foregoing discussion, any one or various combinations of two sensors discussed in connection with a TMS can correspond to the first measurement device connected to an expansion valve, and any one or various combination of two sensors can correspond to the second measurement device. In general, as discussed previously, the first measurement device provides information corresponding to a first thermodynamic quantity to the first control device, and the second measurement device provides information corresponding to a second thermodynamic quantity to the second control device, where the first and second thermodynamic quantities are different, and therefore allow the first and second control device to independently control two different system properties (e.g., the vapor quality of the refrigerant fluid and the heat load temperature, respectively).

In some embodiments, one or more of the sensors shown in a TMS are connected directly to an expansion valve. Control devices can be configured to adaptively respond directly to the transmitted signals from the sensors, thereby providing for automatic adjustment of the system's operating parameters. In certain embodiments, the control device can include processing hardware and/or software components that receive transmitted signals from the sensors, optionally perform computational operations, and activate elements of the first and/or second control device to adjust the control device in response to the sensor signals.

In addition, control system 999 is optionally connected to an expansion valve. In embodiments where an expansion valve is implemented as a device controllable via an electrical control signal, control system 999 is configured to transmit suitable control signals to the first and/or second control device to adjust the configuration of these components. In particular, control system 999 is configured to adjust an expansion valve to control the vapor quality of a refrigerant fluid in a TMS.

During operation of a TMS, control system 999 typically receives measurement signals from one or more sensors. The measurements can be received periodically (e.g., at consistent, recurring intervals) or irregularly, depending upon the nature of the measurements and the manner in which the measurement information is used by control system 999. In some embodiments, certain measurements are performed by control system 999 after particular conditions—such as a measured parameter value exceeding or falling below an associated set point value—are reached.

By way of example, Table 1 summarizes various examples of combinations of types of information (e.g., system properties and thermodynamic quantities) that can be measured by the sensors of system and transmitted to control system 999, to allow control system 999 to generate and transmit suitable control signals to expansion valve 114 and/or other control devices. The types of information shown in Table 1 can generally be measured using any suitable device (including combination of one or more of the sensors discussed herein) to provide measurement information to control system 999.

TABLE 1 Measurement Information Used to Adjust First Control Device FCM Evap Press Press Rec Evap Evap HL Drop Drop Pres VQ SH VQ P/T Temp Measure- FCM x x ment Press Informa- Drop tion Evap x x Used to Press Adjust Drop Second Rec x x Control Press Device VQ x x SH x x Evap x x VQ Evap x x x x x x x P/T HL x x x x x x x Temp FCM Press Drop = refrigerant fluid pressure drop across first control device Evap Press Drop = refrigerant fluid pressure drop across evaporator Rec Press = refrigerant fluid pressure in receiver 110 VQ = vapor quality of refrigerant fluid SH = superheat of refrigerant fluid Evap VQ = vapor quality of refrigerant fluid at evaporator outlet Evap P/T = evaporation pressure or temperature HL Temp = heat load temperature

For example, in some embodiments, expansion valve 114 is adjusted (e.g., automatically or by control system 999) based on a measurement of the evaporation pressure (p_(e)) of the refrigerant fluid and/or a measurement of the evaporation temperature of the refrigerant fluid. In certain embodiments, expansion valve 114 is adjusted (e.g., automatically or by control system 999) based on a measurement of the temperature of a heat load.

To adjust any of the control devices, the compressor, or the pump based on a particular value of a measured system parameter value, control system 999 compares the measured value to a set point value (or threshold value) for the system parameter. Certain set point values represent a maximum allowable value of a system parameter, and if the measured value is equal to the set point value (or differs from the set point value by 10% or less (e.g., 5% or less, 3% or less, 1% or less) of the set point value), control system 999 adjusts an expansion valve to adjust the operating state of the system, and reduce the system parameter value.

Certain set point values represent a minimum allowable value of a system parameter, and if the measured value is equal to the set point value (or differs from the set point value by 10% or less (e.g., 5% or less, 3% or less, 1% or less) of the set point value), control system 999 adjusts an expansion valve, etc. to adjust the operating state of the system, and increase the system parameter value.

Some set point values represent “target” values of system parameters. For such system parameters, if the measured parameter value differs from the set point value by 1% or more (e.g., 3% or more, 5% or more, 10% or more, 20% or more), control system 999 adjusts expansion device 18, etc. to adjust the operating state of the system, so that the system parameter value more closely matches the set point value.

Measured parameter values are assessed in relative terms based on set point values (i.e., as a percentage of set point values). Alternatively, in some embodiments, measured parameter values can be accessed in absolute terms. For example, if a measured system parameter value differs from a set point value by more than a certain amount (e.g., by 1 degree C. or more, 2 degrees C. or more, 3 degrees C. or more, 4 degrees C. or more, 5 degrees C. or more), then control system 999 adjusts expansion device 18, etc. to adjust the operating state of the system, so that the measured system parameter value more closely matches the set point value.

In the foregoing examples, measured parameter values are assessed in relative terms based on set point values (i.e., as a percentage of set point values). Alternatively, in some embodiments, measured parameter values can be in absolute terms. For example, if a measured system parameter value differs from a set point value by more than a certain amount (e.g., by 1 degree C. or more, 2 degrees C. or more, 3 degrees C. or more, 4 degrees C. or more, 5 degrees C. or more), then control system 999 adjusts expansion device 18, etc. to adjust the operating state of a TMS, so that the measured system parameter value more closely matches the set point value.

In certain embodiments, refrigerant fluid emerging from cold plate 116 can be used to cool one or more additional heat loads. In addition, systems can include a second heat load connected to a heat exchanger. A variety of mechanical connections can be used to attach second heat load to heat exchanger, including (but not limited to) brazing, clamping, welding, and any of the other connection types discussed herein.

Heat exchanger includes one or more flow channels through which high vapor quality refrigerant fluid flows after leaving cold plate 116. During operation, as the refrigerant fluid vapor phases through the flow channels, it absorbs heat energy from a second heat load, cooling second heat load. Typically, a second heat load is not as sensitive as a high heat load to fluctuations in temperature. Accordingly, while second heat load is generally not cooled as precisely relative to a particular temperature set point value as heat load, the refrigerant fluid vapor provides cooling that adequately matches the temperature constraints for a second heat load.

In general, the systems disclosed herein can include more than one (e.g., two or more, three or more, four or more, five or more, or even more) heat loads in addition to heat loads depicted. Each of the additional heat loads can have an associated heat exchanger; in some embodiments, multiple additional heat loads are connected to a single heat exchanger, and in certain embodiments, each additional heat load has its own heat exchanger. Moreover, each of the additional heat loads can be cooled by the superheated refrigerant fluid vapor after a heat exchanger attached to the second load or cooled by the high vapor quality fluid stream that emerges from cold plate 116.

The vapor quality of the refrigerant fluid, after passing through cold plate 116, can be controlled either directly or indirectly with respect to a vapor quality set point by control system 999. In some embodiments, a TMS includes a vapor quality sensor that provides a direct measurement of vapor quality which is transmitted to control system 999. Control system 999 adjusts control device depending on configuration to control the vapor quality relative to the vapor quality set point value.

In certain embodiments, a TMS includes a sensor that measures superheat and indirectly, vapor quality. For example, a combination of temperature and pressure sensors measure the refrigerant fluid superheat downstream from a second heat load, and transmit the measurements to control system 999. Control system 999 adjusts the control device according to the configuration based on the measured superheat relative to a superheat set point value. By doing so, control system 999 indirectly adjusts the vapor quality of the refrigerant fluid emerging from cold plate 116.

As the two refrigerant fluid streams flow in opposite directions within recuperative heat exchanger, heat is transferred from the refrigerant fluid emerging from cold plate 116 to the refrigerant fluid entering expansion device 18. Heat transfer between the refrigerant fluid streams can have a number of advantages. For example, recuperative heat transfer can increase the refrigeration effect in cold plate 116, reducing the refrigerant mass transfer rate implemented to handle the heat load presented by high heat load 118. Further, by reducing the refrigerant mass transfer rate through cold plate 116, the amount of refrigerant used to provide cooling duty in a given period of time is reduced. As a result, for a given initial quantity of refrigerant fluid introduced into receiver 110, the operational time over which the system can operate before an additional refrigerant fluid charge is needed can be extended. Alternatively, for the system to effectively cool high heat load 118 for a given period of time, a smaller initial charge of refrigerant fluid into receiver 110 can be used.

Because the liquid and vapor phases of the two-phase mixture of refrigerant fluid generated following expansion of the refrigerant fluid in expansion valve 114 can be used for different cooling applications, in some embodiments, the system can include a phase separator to separate the liquid and vapor phases into separate refrigerant streams that follow different flow paths within a TMS.

Further, eliminating (or nearly eliminating) the refrigerant vapor from the refrigerant fluid stream entering cold plate 116 can help to reduce the cross-section of the evaporator and improve film boiling in the refrigerant channels. In film boiling, the liquid phase (in the form of a film) is physically separated from the walls of the refrigerant channels by a layer of refrigerant vapor, leading to poor thermal contact and heat transfer between the refrigerant liquid and the refrigerant channels. Reducing film boiling improves the efficiency of heat transfer and the cooling performance of cold plate 116.

In addition, by eliminating (or nearly eliminating) the refrigerant vapor from the refrigerant fluid stream entering cold plate 116, distribution of the liquid refrigerant within the channels of cold plate 116 can be made easier. In certain embodiments, vapor present in the refrigerant channels of cold plate 116 can oppose the flow of liquid refrigerant into the channels. Diverting the vapor phase of the refrigerant fluid before the fluid enters cold plate 116 can help to reduce this difficulty.

In addition to phase separator, or as an alternative to phase separator, in some embodiments the systems disclosed herein can include a phase separator downstream from cold plate 116. Such a configuration can be used when the refrigerant fluid emerging from evaporator is not entirely in the vapor phase, and still includes liquid refrigerant fluid.

The foregoing examples of thermal management systems illustrate a number of features that can be included in any of the systems within the scope of this disclosure. In addition, a variety of other features can be present in such systems.

In certain embodiments, refrigerant fluid that is discharged from cold plate 116 and passes through conduit can be directly discharged as exhaust from conduit without further treatment. Direct discharge provides a convenient and straightforward method for handling spent refrigerant, and has the added advantage that over time the overall weight of the system is reduced due to the loss of refrigerant fluid. For systems that are mounted to small vehicles or are otherwise mobile, this reduction in weight can be important.

In some embodiments, however, refrigerant fluid vapor can be further processed before it is discharged. Further processing may be desirable depending upon the nature of the refrigerant fluid that is used, as direct discharge of unprocessed refrigerant fluid vapor may be hazardous to humans and/or may be deleterious to mechanical and/or electronic devices in the vicinity of a TMS. For example, the unprocessed refrigerant fluid vapor may be flammable or toxic, or may corrode metallic device components. In situations such as these, additional processing of the refrigerant fluid vapor may be desirable.

In general, refrigerant processing apparatus can be implemented in various ways. In some embodiments, refrigerant processing apparatus is a chemical scrubber or water-based scrubber. Within apparatus, the refrigerant fluid is exposed to one or more chemical agents that treat the refrigerant fluid vapor to reduce its deleterious properties. For example, where the refrigerant fluid vapor is basic (e.g., ammonia) or acidic, the refrigerant fluid vapor can be exposed to one or more chemical agents that neutralize the vapor and yield a less basic or acidic product that can be collected for disposal or discharged from apparatus.

As another example, where the refrigerant fluid vapor is highly chemically reactive, the refrigerant fluid vapor can be exposed to one or more chemical agents that oxidize, reduce, or otherwise react with the refrigerant fluid vapor to yield a less reactive product that can be collected for disposal or discharged from apparatus.

In certain embodiments, refrigerant processing apparatus can be implemented as an adsorptive sink for the refrigerant fluid. Apparatus can include, for example, an adsorbent material bed that binds particles of the refrigerant fluid vapor, trapping the refrigerant fluid within apparatus and preventing discharge. The adsorptive process can sequester the refrigerant fluid particles within the adsorbent material bed, which can then be removed from apparatus and sent for disposal.

In some embodiments, where the refrigerant fluid is flammable, refrigerant processing apparatus can be implemented as an incinerator. Incoming refrigerant fluid vapor can be mixed with oxygen or another oxidizing agent and ignited to combust the refrigerant fluid. The combustion products can be discharged from the incinerator or collected (e.g., via an adsorbent material bed) for later disposal.

As an alternative, refrigerant processing apparatus can also be implemented as a combustor of an engine or another mechanical power-generating device. Refrigerant fluid vapor from conduit can be mixed with oxygen, for example, and combusted in a piston-based engine or turbine to perform mechanical work, such as providing drive power for a vehicle or driving a generator to produce electricity. In certain embodiments, the generated electricity can be used to provide electrical operating power for one or more devices, including high heat load 118. For example, high heat load 118 can include one or more electronic devices that are powered, at least in part, by electrical energy generated from combustion of refrigerant fluid vapor in refrigerant processing apparatus.

The thermal management systems disclosed herein can optionally include a phase separator upstream from the refrigerant processing apparatus.

Particularly during start-up of a TMS disclosed herein, liquid refrigerant may be present in conduits because the systems generally begin operation before high heat load and/or low heat load is activated. Accordingly, phase separator functions in a manner similar to phase separators to separate liquid refrigerant fluid from refrigerant vapor. The separated liquid refrigerant fluid can be re-directed to another portion of the system, or retained within phase separator until it is converted to refrigerant vapor. By using phase separator, liquid refrigerant fluid can be prevented from entering refrigerant processing apparatus.

In some embodiments, the refrigeration systems disclosed herein can be combined with power systems to form integrated power and thermal systems, in which certain components of the integrated systems are responsible for providing refrigeration functions and certain components of the integrated systems are responsible for generating operating power.

FIG. 12 shows an integrated power and TMS that includes many features similar to those discussed above (e.g., see FIGS. 1 and 8-10 ) with only aspects of the open circuit portion (in this case OCRS 5) shown. In addition, the TMS includes an engine 1200 with an inlet that receives the stream of waste refrigerant fluid. Engine 1200 can combust the waste refrigerant fluid directly, or alternatively, can mix the waste refrigerant fluid with one or more additives (such as oxidizers) before combustion. Where ammonia is used as the refrigerant fluid in OCRS 5, suitable engine configurations for both direct ammonia combustion as fuel, and combustion of ammonia mixed with other additives, can be implemented. In general, combustion of ammonia improves the efficiency of power generation by the engine.

The energy released from combustion of the refrigerant fluid can be used by engine 1200 to generate electrical power, e.g., by using the energy to drive a generator. The electrical power can be delivered via electrical connection to high heat load 118 to provide operating power for the load. For example, in certain embodiments, high heat loads 118 include one or more electrical circuits and/or electronic devices, and engine 1200 provides operating power to the circuits/devices via combustion of refrigerant fluid. Byproducts 1202 of the combustion process can be discharged from engine 1200 via exhaust conduit, as shown in FIG. 12 .

Various types of engines and power-generating devices can be implemented as engine 1200 in this TMS. In some embodiments, for example, engine 1200 is a conventional four-cycle piston-based engine, and the waste refrigerant fluid is introduced into a combustor of the engine. In certain embodiments, engine 1200 is a gas turbine engine, and the waste refrigerant fluid is introduced via the engine inlet to the afterburner of the gas turbine engine. As discussed above, in some embodiments, the TMS can include a phase separator (not shown) positioned upstream from engine 1200. Phase separator functions to prevent liquid refrigerant fluid from entering engine 1200, which may reduce the efficiency of electrical power generation by engine 1200.

In certain embodiments, the thermal management systems disclosed herein operate differently at, and immediately following, system start-up compared to the manner in which the systems operate after an extended running period. Upon start-up, the compressor 104 and a device (usually a fan) moving a cooling fluid (usually ambient air) through the condenser 106 are powered. The compressor 104 discharges compressed refrigerant into the condenser 106. The refrigerant is condensed and subcooled in the condenser 106. Liquid refrigerant fluid enters receiver 110 at a pressure and temperature generated by operation of the compressor 104 and condenser 106.

The thermal management systems and methods disclosed herein can be implemented as part of (or in conjunction with) directed energy systems such as high energy laser systems. Due to their nature, directed energy systems typically present a number of cooling challenges, including certain heat loads for which temperatures are maintained during operation within a relatively narrow range.

FIG. 13 shows one example of a directed energy system, specifically, a high energy laser system 1300. System 1300 includes a bank of one or more laser diodes 1302 and an amplifier 1304 connected to a power source 1306. During operation, laser diodes 1302 generate an output radiation beam 1308 that is amplified by amplifier 1304, and directed as output beam 1310 onto a target. Generation of high energy output beams can result in the production of significant quantities of heat. Certain laser diodes, however, are relatively temperature sensitive, and the operating temperature of such diodes is regulated within a relatively narrow range of temperatures to ensure efficient operation and avoid thermal damage. Amplifiers are also temperature-sensitive, although typically less sensitive than diodes.

To regulate the temperatures of various components of directed energy systems such as diodes 1302 and amplifier 1304, such systems can include components and features of the thermal management systems disclosed herein. In FIG. 13 , cold plate 116 is coupled to diodes 1302 as well as to amplifier 1304. Alternatively, a first cold plate 116 can cool both diodes 1302 and amplifier 1304 or one of diodes 1302 and amplifier 1304, with a second cold plate 116 (or recuperative heat exchanger 122) cooling the other of diodes 1302 and amplifier 1304. The other components of the thermal management systems disclosed herein are not shown for clarity. However, it should be understood that any of the features and components discussed above can optionally be included in directed energy systems. Diodes 1302, due to their temperature-sensitive nature, effectively function as a high heat load 118 in system 1300, while amplifier 1304 functions as a low heat load 120.

System 1300 is one example of a directed energy system that can include various features and components of the thermal management systems and methods described herein. However, it should be appreciated that the thermal management systems and methods are general in nature, and can be applied to cool a variety of different heat loads under a wide range of operating conditions.

A number of embodiments have been described. Nevertheless, it will be understood that various modifications may be made without departing from the spirit and scope of the disclosure. Accordingly, other embodiments are within the scope of the following claims. 

What is claimed is:
 1. An evaporator, comprising: a cold plate configured to extract heat from one or more heat loads in proximity to the evaporator, the cold plate comprising: a housing, and a plurality of channels disposed through the housing, with at least one of the plurality of channels being a meandered channel.
 2. The evaporator of claim 1, wherein the meandered channel has at least one curved region in the meandered channel.
 3. The evaporator of claim 1, wherein the meandered channel has one curved region in the meandered channel and the meandered channel has a net internal angle greater than zero degrees and up to 360 degrees.
 4. The evaporator of claim 1, wherein the meandered channel has more than one curved region in the meandered channel and the meandered channel has a net internal angle between and including zero degrees up to but not including 360 degrees.
 5. The evaporator of claim 2, wherein the at least one curved region in the meandered channel is defined as a first section of the channel that forms a non-zero degree angle along the length of the channel with respect to a second section of the channel.
 6. The evaporator of claim 1, wherein the cold plate is a first cold plate and the evaporator comprises a plurality of cold plates including the first cold plate.
 7. The evaporator of claim 6, wherein each of the plurality of cold plates includes at least one meandered channel.
 8. The evaporator of claim 6, wherein at least one of the plurality of cold plates includes the at least one meandered channel.
 9. The evaporator of claim 6, further comprising: a plurality of inlet headers fluidly coupled to inlets of the plurality of cold plates; and a plurality of exit headers fluidly coupled to outlets of the plurality of cold plates.
 10. The evaporator of claim 9, further comprising an inlet distributer, comprising: an inlet configured to receive refrigerant fluid; and a plurality of outlets, each of the plurality of outlets fluidly coupled to a corresponding one of the plurality of inlet headers.
 11. The evaporator of claim 9, further comprising: an exit collector having a plurality of inlets configured to receive refrigerant fluid from the plurality of exit headers.
 12. A thermal management system, comprising: an evaporator that comprises at least one cold plate configured to extract heat from one or more heat loads in proximity to the evaporator, the cold plate comprising a housing and a plurality of channels disposed through the housing, with at least one of the plurality of channels being a meandered channel; a receiver configured to store refrigerant fluid, the receiver disposed in a refrigerant fluid path with the evaporator; and an expansion valve positioned between the receiver and the evaporator in the refrigerant fluid path, the expansion valve configured to expand the refrigerant fluid from the receiver and deliver a mixed liquid/vapor refrigerant to the evaporator.
 13. The thermal management system of claim 12, wherein the meandered channel comprises at least one curved region in the meandered channel.
 14. The thermal management system of claim 12, wherein the meandered channel comprises one curved region in the meandered channel and the meandered channel has a net internal angle greater than zero degrees up to 360 degrees.
 15. The thermal management system of claim 13, wherein the at least one curved region in the meandered channel is defined as a first section of the channel that forms a non-zero degree angle along the length of the channel with respect to a second section of the channel.
 16. The thermal management system of claim 12, wherein the cold plate is a first cold plate and the evaporator comprises a plurality of cold plates including the first cold plate and each of the plurality of cold plates includes at least one meandered channel.
 17. The thermal management system of claim 12, wherein the evaporator further comprises: a plurality of inlet headers coupled to inlets of the plurality of cold plates; and a plurality of exit headers, coupled to outlets of the plurality of cold plates.
 18. The thermal management system of claim 16, wherein the evaporator further comprises: an inlet distributer comprising an inlet configured to receive the refrigerant fluid and a plurality of outlets, each of the plurality of outlets fluidly coupled to a corresponding one of the plurality of inlet headers; and an exit collector comprising a plurality of inlets configured to receive the refrigerant fluid from the plurality of exit headers.
 19. The thermal management system of claim 12, wherein the refrigerant fluid path is an open-circuit refrigerant fluid path, the system further comprising: a back-pressure regulator coupled to an outlet of the evaporator and configured to control vapor pressure upstream of the back-pressure regulator.
 20. The thermal management system of claim 18, wherein the refrigerant fluid path further includes a closed-circuit refrigerant fluid path, and the system further comprises: a compressor configured to receive refrigerant vapor from the evaporator and compress the refrigerant vapor; and a condenser configured to condense the compressed refrigerant vapor from the compressor.
 21. The thermal management system of claim 12, wherein the expansion valve is configured to perform a constant-enthalpy expansion of the liquid refrigerant fluid to generate the refrigerant fluid mixture.
 22. The thermal management system of claim 12, wherein the refrigerant fluid comprises ammonia.
 23. A method of operating an evaporator, comprising: transporting refrigerant fluid through a plurality of channels in a cold plate according to a required mass velocity range and a required mass flow rate demand, with at least one channel of the plurality of channels being a meandered channel; transferring refrigerant through the plurality of channels to absorb an amount of heat from a heat load to provide at an evaporator outlet a preset exit vapor quality; and maintaining a heat load temperature band of the heat load, while thermally conducting the heat from the heat load to the plurality of refrigerant channels within the cold plate.
 24. The method of claim 23, wherein the meandered channel has at least one curved region in the meandered channel.
 25. The method of claim 23, wherein the meandered channel has one curved region in the meandered channel and the meandered channel has a net internal angle greater than zero degrees up to 360 degrees.
 26. The method of claim 24, wherein the at least one curved region in the meandered channel is defined as a first section of the channel that forms a non-zero degree angle along the length of the channel with respect to a second section of the channel. 